Resiliently mounted fluid bearing assembly

ABSTRACT

A fluid bearing assembly, e.g., air or gas, of the hydrostatic or hydrodynamic type, for use as a spindle and the like, includes spaced journals mounted on a shaft for rotation therewith. The housing includes spaced support bearings cooperating with the journals to form a gas gap therebetween. The support bearings are resiliently mounted in the housing for axial, radial and angular movement through the use of continuous resilient support elements positioned between the housing and the support bearings and continuously in contact with each. The resilient support elements are preferably self-energizing and comprised of relatively low coefficient of friction materials to reduce the friction of the system thus allowing rapid adjustment of the bearing in response to loads on the shaft. Various details are described with respect to the various structures illustrated, including the details of an air bearing spindle assembly of relatively light weight and capable of operating at speeds in excess of 100,000 RPM.

This is a division of application Ser. No. 035,627, filed Apr. 3, 1987,now U.S. Pat. No. 4,828,403.

FIELD OF THE INVENTION

The present invention relates to a fluid bearing assembly and moreparticularly to an improved and relatively simple resiliently mountedfluid and gas bearing assembly including a rotor or shaft membersupported by at least two spaced resiliently mounted bearings each ofwhich is movable, for example, in essentially three axes, radially,axially and angularly with respect to a support structure therebyassuring a proper gap between the shaft and the support bearing,especially in response to thrust loads, thus providing a precisionassembly of relatively simple structure which is relatively easy andinexpensive to manufacture and of relatively light weight.

BACKGROUND OF THE PRIOR ART

Fluid bearings, especially gas bearings, are known in the art and whichare effective for the purposes intended. For example my prior U.S. Pat.Nos. 3,249,390 and 3,476,451, whose disclosures are incorporated hereinby reference, describe fluid bearing systems and gas bearing systemswhich operated satisfactorily but which were difficult and relativelyexpensive to manufacture and which had a relatively large profile andwere relatively heavy due to the large profile.

Those structures includes spaced bearings and a rotor or shaft in whichthe forward bearing was relatively fixed, during operation, with respectto axial and radial and angular motion. In one form, there was no springsupport for one bearing while the other was supported at spaced pointsalong the outer periphery. In another form, the hydrodynamic form, theforward support bearing was fixed axially. For the purposes of thisinvention, the terms shaft and rotor refer to the rotating part whilethe term bearing refers to the fixed, i.e., non-rotating element whichsupports the shaft for rotation on a fluid or gas film. While the rearbearing of these prior structures permitted limited axial movement, andlimited radial and angular movement, the assembly as a whole was limitedin movement in that only one end of the rotor or shaft was capable oflimited movement. As a result, in the event of a significant forceaxially of the rotor or shaft, or angularly or radially of the moveableend of the rotor, the forward end of the rotor would tend to contact theforward supporting bearing and bottom out on that bearing. The resultwas a catastrophic failure of the bearing assembly because of thefailure to mount the supporting bearing in a resilient manner, to bedescribed.

Thus, there are limitations in that prior art structure as described inmy prior patents. More specifically, the support bearings were supportedat three separate points disposed 120 degrees to each other such thatthe forward support bearing was essentially fixedly supported inoperation, but with relatively severely limited movement. The forwardsupport, during operation, was urged axially of the rotor to seatagainst the opposed face of the supporting housing by annular shoulderson the support. In order to move radially, it was necessary for theforward bearing to overcome the friction of the shoulders against thehousing. Seals between the support bearing and the housing were in theform of piston rings, which were split, somewhat similar to those ofautomotive piston rings. Due to the three point support and the factthat the support bearing was seated against the housing, the forwardsupport bearing was not capable of angular movement with respect to therotor axis and to the extent that there was any movement at all, therewas relatively high friction.

As a result the friction, coulomb friction, of the forward supportbearing assembly, the friction of the total bearing assembly wasrelatively high and was used as a dampening effect. It was believed thenthat the coulomb friction, i.e., the friction which has to be overcometo effect relative movement of non-lubricated parts, was operative toprovide a dampening effect between the forward support bearing and theremaining structure to overcome the radial oscillatory motion of theassembly.

The total friction of the assembly, however, was relatively high,although it was thought at that time to be acceptable. The system wasintended to be a two degree freedom of movement system including twosprings and two masses. The rotor was one mass and the support bearingswere the second mass while one spring was the gas film and the otherspring was the radially located separate three spring system, asdescribed, between the support bearing and the housing. Since one of thesupport bearings was fixed axially in operation and capable of someradial movement at a relatively high friction, the effectiveness of thesystem as a two degree of freedom system was somewhat less than desired.This became apparent when one attempted to use the prior describedbearing at relatively high rotational speeds and where axial loads wereplaced on the rotor.

In such a case, the entire structure must be capable of respondingrelatively rapidly to changes in load and pressure in order to operatesatisfactorily. To achieve this, it was determined that axial, angularand radial motion was desirable in order to provide the relatively rapidresponse needed to accommodate changes in load and pressure, andespecially in response to thrust loads. Further, such motions had to beachieved with relatively low friction, i.e., relatively low coulombfriction, so that the parts responded quickly and relatively smoothlyrather than being relatively non-moveable until the relatively highcoloumb friction was overcome.

In addition to the above, it is known in the prior art to use O-rings asa seal between the bearings and the support housing in which thebearings are mounted. The difficulty with O-rings is that they areusually compressed radially to function as a seal and are of limitedresiliency in a radial direction and tend to perform erratically as thetemperature increases. Further, O-rings tend to degrade over time due tochemical changes in the material of which the rings are made. Normally,these O-rings are not considered to be resilient support elements, butmore of a sealing element with the result that they tend to berelatively high friction elements, even if coated with low frictionmaterial such as polytetrafluorethylene or other fluorocarbon or lowcoefficient of friction materials. As such, the inherent design of theO-ring is somewhat controlled in the sense that one has to design theenvironment in which they are used to match the structure to thecharacter of the O-ring. This imposes some sever limitations on the useof O-rings, especially if it is necessary to provide a resilient supportoperative over a wide range of operating temperatures, for relativelylong periods of time, and which is capable of adjusting to differenttypes of relative movement between the mating parts.

In general, gas lubricated bearings are normally assumed to have a givenspring constant and a small amount of inherent damping. The springconstant may be measured by plotting a force displacement curve whilethe damping coefficient may be determined by ascertaining the decreasein amplitude as a function of time of a bearing system after it issubjected to an impact load.

However, gas bearings are different from more conventional bearings inthe sense that there is a finite time lag between the initialapplication of a displacement force and the time required for thebearing to reach a steady state condition. Upon extensive investigationof experimental evidence, it was noted that this time constant or timelag is difficult to measure because it is obscured by the bearing'sdamping. This time constant can be calculated and becomes larger ifpocket type bearings are used. It can also be shown by computersimulation that, in general, the longer the time constant (first-orderlag), the more unstable the system becomes. This time lag is caused bythe fact that a finite amount of time is required for the gas to flowout of the bearing or into the bearing until a new pressure distributionhas been reached.

It is known that a pocket type bearing has a greater restoring force,i.e., the force tending to establish a balancing equilibrium condition,and since a resiliently mounted fluid bearing is more stable, a bearingsystem employing pockets may be utilized. Thus, a resilient mountedsystem may be designed for a gas bearing in which the added advantage ofthe greater restoring force associated with pocket type bearings can beused even though they have a long time constant.

As pointed out in the patents previously referred to, there areapplications in which the use of conventional lubricants in a bearingpresents significant problems. For, example, in cryogenic applications,ordinary petro-lubricants and the like, or silicone lubricants becomeunusable due to the fact that they become viscous or solid. In oxidizingenvironments, some lubricants are totally unacceptable due to possiblecatastrophic failure. Materials such as graphite may be used, but are oflimited utility. All of these, and other factors are discussed in theprior patents referred to previously.

In instances in which the gas bearing is used to support high speedspindles, in the range of 15,000 and preferably 80,000 to 130,000 orhigher RPM, such as those used in circuit board drilling machines orother high RPM accurate drilling equipment, the debris caused by suchoperations produces "cuttings" which may be quite abrasive or otherwisehostile. Properly designed in accordance with this invention, rotationalspeeds as high as 450,000 RPM may be achieved. The cuttings usually arerelatively light weight and are in the form of relatively fine dustwhich may infiltrate the bearing or the spindle and which may shortenthe life thereof because of the introduction of relatively abrasiveparticles into the relatively small space between normally precision andmanufactured parts which rotate relative to one another. For example,roller bearing spindles suffer the problem of a relatively high wear dueto the presence of fine abrasive particles unless some special andsometimes relatively heavy and expensive structures are provided toprevent entry of such abrasive particles into the working space betweenthe relatively stationary and rotating elements of the bearing andspindle assembly. Such debris may cause serious damage to O-ring sealsystems.

In the case of that type of environment, the use of a gas lubricatedbearing, in accordance with the present invention offers singularadvantages because the bearing structure is relatively immune from suchrelatively hostile environments since the exhaust of the gas used in thebearing prevents most of the debris from entering the bearing or spindlehousing. Nonetheless, there are other concerns regarding the use of gasbearings due to some of the problems traditionally associated therewith,some of which have already been discussed.

Gas lubricated bearings and spindle assemblies are known in the art andsome of these have attempted to overcome the problems noted. However,these structures, some of which are described in patents noted and someof which are currently commercial products, such as the gas bearingavailable from Federal Mogul Corporation under the trademark WESTWIND,have not solved the problem of a light weight and reliable structurecapable of operating at comparatively high rotational speeds inenvironments which generate hostile and abrasive products. For example,the Westwind ball bearing system weighs about five or more pounds, whilethe Westwind gas bearing system is heavier and requires air filters forthe incoming air and air driers to reduce moisture thereof as well asrefrigeration systems to cool the incoming air.

One of the advantages of the present invention is he provision of aresiliently mounted gas lubricated bearing which is essentially immuneto such a hostile environment since the exhaust gas discharged by thebearing prevents essentially all of the debris from entering the bearingor spindle housing. The obvious advantage is that there is no need forrelatively heave sealed and grease lubricated assemblies, whileproviding a relatively light weight structure capable of operation atcomparatively high speeds and in environments in which corrosiveproducts tend to adversely affect the performance of even sealedlubricated assemblies.

More recently, efforts have been made to overcome the above notedproblems through the use of conventionally lubricated bearings. Theresults have been products of relatively low useful life and which arerelatively heavy and bulky in order to provide the encapsulationshielding needed to reduce the effects of such corrosive and harmfulproducts. Even so, the results have been marginal at best as compared tothis invention.

In the case of gas lubricated bearings in general, lubrication isachieved by containing a thin film (about 0.0005 of an inch) of gasbetween an accurately machined shaft journal and the bearing. The resulthas been to provide bearings for special applications which aresubstantially insensitive to super extremes in temperature. Further,since there is no contact between the relatively rotating parts, thereis no coulomb friction, heating or rubbing, and consequently no wearfrom such effects. As is understood, properly designed gas lubricatedbearings may be operated at very high rotational speeds, with long lifeand stable performance characteristics. Nonetheless, the prior designshave yet to reach that level of performance in terms of light weight andhigh rotational speeds which gas lubricated bearings are capable ofachieving.

Gas lubricated bearings may be of the hydrostatic or hydrodynamic type.Both types are contemplated by the present invention. In a hydrostaticbearing, gas is continuously supplied to the bearing interspace under apredetermined pressure. In hydrodynamic gas bearings, the gaslubricating film is self-maintaining when relatively high tangentialvelocities are reached, relative to the bearing surface and shaft andmay, if desired, by totally isolated from other sources of gas. Thiscapability to be sealed off causes hydrodynamic bearings to beattractive in applications such as gyros and the like where themaintenance or use of a source of pressurized gas may not be feasible.

Yet, the machining tolerances for bearings of the hydrodynamic type aresignificantly more stringent than for the hydrostatic type of bearing.In the hydrodynamic bearing, the bearing interspace gap is usually lessthan about 0.0001 of an inch. This relatively small gap gives rise todetectable viscous friction at high bearing speeds and is manifested asdrag on the rotary shaft. Further, the self-maintaining of the filmpressure inherently limits the versatility of the load capability andthe selection of critical (resonant) angular speeds of the bearing.

In the case of hydrostatic bearings the machining tolerances are relaxedsomewhat as compared to hydrodynamic bearings but are nonethelessstringent. Moreover, the gas feed into the bearing interspace must, insome cases, be angularly symmetric in order to support the rotary shaftthus to preclude imbalance with respect thereto. Even so, thehydrostatic approach provides a more versatile and stable bearingbecause of its control ability and larger spacings.

The prior art devices in the gas bearing area, including the patentspreviously referred to, achieve support for axial and radial thrusts orloads by the use of a pair of juxtaposed spaced cylindrical surfaces anda pair of juxtaposed spaced radially arranged surfaces wherein onesurface of each pair is on the shaft and its juxtaposed counterpart isor is supported on a stationary frame or housing. There is noappreciable cooperation between the two bearing parts, their supportingforces being mutually orthogonal.

A generic deficiency of the prior art gas bearing structures, includingthose of the patents identified, is that regardless of the machiningtolerances and regardless of the care with which the shaft is loaded inuse, a finite rotary imbalance exists which causes an oscillation inconjunction with the elastic restoring force of the supporting gas film.While these bearing assemblies are thought of as being two degree offreedom systems composes of two springs and two masses in that the rotorand bearing are the masses and the gas film and supposedly resilientmounting are the spring system, they do not behave as two degree offreedom systems.

In practice, the supporting gas film has a relatively low "spring"constant and the resonant critical speed of the shaft is so low as to bea severe limitation on high rotary speed applications of gas lubricatedbearings. The severity of the resonance problem is caused by the factthat the near zero viscous friction of the gas film affords near zerodamping of the oscillating bearing. Consequently, the bearing structuretends to oscillate without limit until the moveable element strikes thestationary bearing or bushing. Typically the resulting coulomb frictionprecludes driving the shaft above the critical frequency or causesdestructive wear or both. Where the coulomb friction is relatively high,as is the case in the structures of the patents referred to, theresponse of the bearing system as a whole to changes in load, pressureand the like is relatively slow and this tends to promote bearing"crashes" or limits the use of the bearing in terms of rotational speed.In effect the system is essentially fixed and relatively incapable ofaxial, radial and angular movement to compensate effectively for suchvariations during operation of the bearing system.

Another difficulty with the gas bearings of the prior art is loss of gasfilm pressure with the result that there is metal-to-metal contact atcertain rotational speeds. It is believed that this is due to "whirl", aphenomenon which is caused by the rotating shaft being displaced, due toits weight, off-center with respect to the axis of the stationarybushing. In effect, the shaft rotates on an axis such that one end ofthe shaft circumscribes a circular track, with the result that the shaftis closer to the bushing at one point than at others, and the shaftexperiences angularly out-of-balance viscous drag. This drag, or thereaction which it produces, is usually in the of the shaft rotation andeffectively causes a rotation of the shaft which is effectivelyangularly unsymmetrical. The whirl rotation, being due to thereverse-directed drag, is usually the same as the shaft rotation and isseen by the gas support film as a reduction in shaft velocity. Thus,with increases in the whirl velocity, since it is in the same directionof shaft rotation, a significant reduction in the support capability ofthe fluid film may occur with the result that there is metal-to-metalcontact between the rotating and non-rotating parts at high speeds.While the restoring force and the whirl resistance of the film may beinfluenced, to some degree, by the fluid pressure in the case of ahydrostatic bearing, and the mass of the shaft may be minimized in orderto increase the critical speed, these represent compromises in thestability and load capability of the bearing and do not represent agenerally applicable solution to the problem.

In general, the solution to the problem, as set forth by this invention,is to make the interior of the bearing structure resilientlyself-adjusting so that it responds rapidly to these events. This may beaccomplished by relative angular, rotational and axial movement of thesupport bearings. By contrast, the structure of the patents previouslyidentified is such that the support bearings are not moveable angularlyat the front end and the bearing has a relatively high friction in aradial direction, with little if any movement in an angular direction.While that system appears to be a true two degree of freedom system, instructure and operation it is not.

Thus, it is an object of the present invention to provide an improvedfluid, preferably a gas, resiliently mounted lubricated bearing andmethod which are not subject to the disadvantages of the prior artsystems referred to above.

Another object of the present invention is to provide a bearingstructure of the type to be described in which the bearing assemblywhich provides the fluid or gas film for rotation of the shaft isresiliently mounted in essentially a free floating condition.

An important object of this invention is the provision of a fluid andpreferably a gas bearing system, capable of operation at relatively highrotational velocities, in which the bearing assembly is flexibly andresiliently mounted to a support structure so that the relative movementof the bearing and the shaft is such that the gap between the fluidsupported shaft and the bearing is maintained to permit the entirestructure to compensate for relative axial movement of the shaft and tocompensate for relative angular and radial movement of the shaft withrespect to the assembly which supports the bearings and the shaft.

It is another object of this invention to provide a gas bearing systemin which a pair of juxtaposed surfaces, resiliently mounted, providesresilient and compensating support for both axial and radial thrust ofthe rotatable shaft.

Still another object of this invention is the provision of a bearingsystem of the type described in which the gas or other fluid filmthickness in both a radial and axial direction is adjustable by theresilient axial or radial relative movement of one of the relativesupporting surfaces through a unique resilient mounting thereof.

Another object of this invention is the provision of a fluid andpreferably gas bearing system in which frictional damping is coupled tothe radially oscillating shaft without solid-to-solid contact betweenthe rotating shaft and the supporting structure by the provision of aunique flexible and resilient mounting which permits relatively rapidadjustment of the relative parts to maintain the required fluid and gasgap between the relatively moving parts.

Yet another object of this invention is the provisions of a bearingsystem of the type described in which the allowable amplitude of radialoscillation of the shaft at critical frequency, without solid-to-solidcontact, may be increased without increasing the quiescent thickness ofthe fluid supporting film through the use of a unique resilientsupporting system.

Still another object of this invention is the provision of an improvedthree-degree of freedom resiliently mounted bearing system of the typedescribed in which the thickness of the fluid, preferably a gas, film isessentially self-adjusting by the provision of an effectivelyself-adjusting film thickness achieved through the use of a uniqueresilient mounting system of the bearing system which allows the weightand profile of the bearing system to be reduced while achievingrelatively high rotational speeds through the achievement of essentiallythe required compensation due to relative axial, angular and radialmovement of the relatively moving parts.

SUMMARY OF THE INVENTION

The above and other objects of this invention are achieved in accordancewith the present invention through the provision of a relatively simpleimproved bearing support system which provides a resilient mounting ofthe bearing support assembly with respect to the support structure suchthat the shaft is supported on a fluid or gas film, which acts as alubricating film, in a true three degree of freedom system.

Thus, in accordance with this invention, the shaft is supported forrotation on a fluid or gas film by spaced resiliently mounted andpreferably spaced bearings. The bearings are resiliently mounted byrelatively low friction support elements preferably having inner andouter low friction support surfaces which may also act as seal surfaces.These resilient support elements are positioned between the bearings anda support structure for the bearings such that the inner low frictionsurface is continuously in contact with and fully surrounds the bearingwhile the outer low friction surface is continuously in contact and iseffectively surrounded by the opposed surface of the supportingstructure. Resiliency is achieved through the use of spring elementsassociated with the low friction elements.

This type of support permits relative axial movement of the shaft andpermits relative angular and radial movement of the bearing with respectto the support such that the fluid gap between the bearing and the shaftis maintained and rapidly adjusted during rotation of the shaft. In thisway, the bearing is allowed motion which is oscillatory with its ownresonant frequency due to its resilient supporting structure and mass.Friction between the resiliently mounted bearing and the supportstructure, which is relatively low in comparison to prior structures,allows relatively rapid response to dampen out the radial oscillatorymotion of the bearing relative to the support structure.

Briefly, in operation, oscillation of the rotary shaft drive theradially "floating" bearing through the supporting gas or fluid film"spring" as a linkage. This permits the shaft a greater amplitude ofoscillation without effectively increasing the thickness of thesupporting film. At the same time, the friction experienced by thesupporting bearing is coupled back through the fluid film linkage to theradially oscillating shaft thereby subtracting from its oscillatoryenergy without solid-to-solid contact. The natural frequency of thesuspended supporting bushing is relatively low so that once the shafthas passed through its critical frequency, there are no furtherresonance problems and the shaft may be rotated at higher speeds withever decreasing oscillation amplitude.

One of the deficiencies of conventional roller bearings is in precisionsurface grinding applications wherein bearing noise is substantiallyalways present and is manifest as imperfections or "noise" on the groundsurface. This occurs because the bearings, typically ball bearings, areinherently imperfect and drive the shaft in a regular imperfectrotation. One of the advantages of this invention is that the bearingoperates above its critical (or resonant) frequency of rotation androtates regularly and essentially perfectly about its true axis ofrevolution. If a grinding wheel affixed to the shaft is dressed at theoperating speed of the fluid supported and lubricated bearings of thisinvention, then the grinding wheel system will rotate essentiallyperfectly about its true and real axes in a manner to permit "noiseless"precision surface grinding.

Still another advantage of the bearing assembly of this invention isthat since the bearing rotates about its true mass, it can effectivelybe utilized for high speed miniature drilling applications. This ispossible because the drill holding mechanism can be manufactured using asuitable and well known grinding system, after balancing, and duringactual rotation of the assembly. This ensures almost perfect alignmentof a miniature drill with the axis of rotation, which would alsocoincide with the center of gravity of the rotating shaft.

As will be apparent, the resilient mounting of the bearings through theuse of self-energized spring biased resilient support elements which mayalso act as seals offers a structure having greater side load capacitydue to forces imposed on the end of the shaft when used in a milling orrouting machine. This is because the flexible and resilient supportelements allow the bearings themselves to tilt somewhat independently.Properly located, as will be described, the tilt axis of the shaft dueto side load on one end and the tilting of the bearing can be made tohave the same angle. The result is that a much greater side load may beapplied on the bearing assembly before rubbing occurs. The design ofthis invention has greater side load capacity than prior art gasbearings.

The feature of the present invention of using self-energizing resilientbearing support elements which may also act as seals provides greatshock capacity by allowing the rotor or shaft to move. Typically theself-energizing resilient support elements are in the form of a V- orU-shaped double lip element having an inner and outer continuous andflexible relatively low coefficient or friction support surface and anassociated spring member. The portion of the resilient support elementforming the lip-like support elements preferably includes a springmember associated with the lip-like support elements. These componentseffectively form a resilient and self-energizing support element for thesupport bearings and these support elements are sufficiently flexible toallow the rotor or shaft to move slightly.

If the spring constant of the resilient support elements and spring andthe thrust take up capacity is somewhat less relative to the gas bearingfluid spring constant, this allows the support bearings to move andtherefore allows the rotor or shaft to move. Due to its own inertialforces, a much greater impact can be withstood by the bearing assemblyof this invention.

Such greater impacts may occur in such operations as drilling with arapid feed rate and a rapid repetition rate which causes an impact typecontact between the drill and the material being drilled. This, in turn,transmits this impact to the rotor/shaft and to the support bearings.

In accordance with this invention, a thrust take-up system is providedin addition to the radial and angular resilient mounting of the supportbearings, preferably in the form of a resilient mounting to take upaxial movement of the shaft in response to axial loads. The thrust-takeup system again involves the use of a resilient support system which mayinclude thrust springs. The main advantages of the thrust springs, whichform part of the resilient support elements, are that they allow forself-adjustment and self-alignment; allow a greater amount of shock tobe absorbed without causing the bearings to bottom out, whiletransmitting the thrust to the support structure.

In accordance with this invention, there are two basic regimes by whichthe springs which form part of the resilient support elements areassociated with respect to the support bearings for performing thefunction of transmitting forces and for holding the various parts in therelative correct position. In the case of relatively small gas or fluidbearing assemblies, it is easier to carry all thrust loads, which arenot large due to the small size of the support bearing, through thestructure of the resilient support elements themselves. If a roundgarter-type spring is used in the resilient support elements, then thatspring has to transmit essentially the entire load. If a flat plate typeof spring is used, then the construction of the support bearings in thearea containing the spring can be made such that essentially all of theaxial loads are transmitted directly through the support surfaces of thelip-like elements, from the support bearings into the supportingstructure.

A structure is also disclosed wherein the resilient support element isretained such that essentially all of the axial load goes directly fromthe support bearing into the spring and into the housing without goingthrough the lip-like support elements. In that particular form, the onlyloads that the lip-like support elements see, other than radial loads,are the gas pressure loads which keep the lip-like support elementssealed to the associated structure.

Since the springs and their spring constants may be controlled, as isknown in the art, forms of the present invention may be provided suchthat there will be greater thrust capacity in one direction than in theother. This is achieved since two springs are used in each direction,one generated by the gas film and the other being that of the mainsprings which take the axial load. Therefore, it is possible inaccordance with the present invention to fabricate the rear main springof a lower spring constant than that of the front main spring. Upon athrust force in a backward direction, the front spring will decrease itsforce at a much more rapid rate than the back axial spring. This allowsa larger gas film gap in the front bearing than the correspondingsmaller gap in the back, so that the net force will be greater than in aconventional system where the gas film is more or less equal to eachend. This added effect of being able to obtain greater thrust in onedirection than the other may be utilized where necessary, such as indrilling applications where greater thrust is required during actualdrilling than during the drill extraction phase.

In the case where one desired great thrust capacity in a small size orincreased thrust capacity in a given size bearing, an embodiment of thisinvention is described whereby three effects are taken advantage of toobtain a much greater thrust capacity in any given size that any othergas bearing system known in the prior art. This added feature is fullydescribed and may be summarized as follows: The area at each end of thesupporting bearings, which are conical and turned outward, are such thatthe larger area is utilized as a thrust piston type of device. Soarranged, the bearing system is maintained in equilibrium when in a nothrust load condition to prevent the shaft from rubbing on surroundingparts and also to align or position the shaft in the desired relativeposition. Upon the application of a thrust load, the conical portion ofthe support bearing generates the reactive force required to counteractthe applied thrust. The thrust is more effective in the direction of theangle of the cones for three reasons. First, the force closes up the gaspressure gap which in itself increases the thrust pressure in the gasfilm area itself. The second effect is to decrease the amount of flowinto the equilibrium volume on the large surface area cone. This, inconjunction with the relatively small amount of movement allows the gasexhaust area at the large end of the cone to be increased thus providinga combination of two effects which reduce the gas pressure acting on aflat end surface of the cone. This pressure is reduced because of asomewhat smaller gas gap film which does not allow as much pressure intothe volume and the increasing exhaust area out of the volume so as toreduce the gas pressure on the end, and therefore causes a reduction ofthrust in the direction from which the thrust load is applied. Thisreduction of thrust therefore increases the thrust counter to thedirection from which the thrust loads are applied. This type of bearingassembly can have almost twice the thrust capacity as compared with aconventional gas bearing of the same projected area. On the conicalsupport bearings, the thrust loads are generated by the projected areain an axial direction, while the radial loads are generated by the areaprojected 90 degrees to the axis.

The bearing system of the present invention may be configured such thatall of the air exhausts out of any given port located almost anywhere onthe unit. In applications that produce dirt, such as high speed drillingof printed circuit boards, all of the air can be made to exhaust out ofthe front of the unit, where the shaft protrudes. The exiting airprevents any of the contaminants from getting into the unit and keepsthe work area clean. It also allows for cooling of the drill by theexhaust airflow across it.

Where a larger bearing capacity is needed, and therefore bearings whichare larger than the armature of the motor which drives the system (therotor being one part and the housing being the stator), it may benecessary to assemble and reassemble the bearing/armature uponinstallation. The bearing system design as herein described may be madeso that the detachable bearing, if it is damaged by exceeding itscapacity, may be replaced relatively simply and inexpensively. Also thesupport bearings may be easily replaced since they are one piece and arerelatively easy to move. Unlike gas bearings which are of a moreintegrated design, the bearing system of this invention is of a modulardesign and only the damaged or worn out parts need be replaced. Themodular design of the bearing system of this invention is achievedthrough the unique suspension of the support bearings on the resilientsupport elements.

In the case of bearing systems in accordance with this invention inwhich the support bearings are conical, there is the advantage that onlytwo surfaces need alignment and the system is self-aligning. Duringoperation at high speeds, the stiffness of the gas film of the bearingsystem of this invention does not have to be as high as in bearingswhere the spring constant of the gas film constitutes the springconstant of the system. For this reason, the very small orificesrequired in a conventional gas bearing are not necessary in the case ofthis invention. The advantage is that it is not necessary to filter theair to the same extent for flow through small orifices.

In the present invention, where the gas film does not have to have ashigh a spring constant, fewer air supply holes of larger diameters maybe used. For example, in most applications, only three or four holes areneeded and these may be of relatively large diameter. For example, inone form of bearing system in accordance with this invention whichsupports a 3/4 pound rotor operating a 125,000 RPM, each support bearinghas four orifices, larger than 4 mils (0.004 of an inch) in diameter andpreferably each in the range of about 0.010 to 0 0.018 of an inch (10 to18 mils) in diameter. These relatively large apertures virtuallypreclude occlusion of the openings by dirt and eliminate the need tofilter incoming air with filters which removes all debris above arelatively small diameter. Further, since the bearing system isself-adjusting, the gap dimension varies with operating conditions andthus, there is no need to have the high degree of air filtration neededwith some of the prior art systems. The bearing system of this inventionis less likely to be damaged by failure of the filtering system.

Since the bearing system of this invention is modularized, differentbearing surface materials may be used and interchanged as needed for aparticular application. For example, for optimum protection of thesupport bearing surfaces, should contact be made during operation, thematerials for the relatively rotating parts may be selected from avariety of materials. Hard anodized aluminum cones operate well withnon-rotating parts such as non-porous graphite composites. High strengthalumina oxide cones operating in conjunction with non-porous graphite orcarbon composite has non-galling and good wear in the event ofaccidental contact between the parts. It will be apparent that othermaterials may be used and that is one of the advantages of thisinvention.

It will be apparent that the advantages discussed and others will bereadily understood by those skilled in the art from the various examplesof structures which illustrate the present invention, all of which arebest understood from a consideration of the following description takenin connection with the drawings which are presented by way ofillustration only and are not to be considered as limiting the presentinvention.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a diagrammatic view, partly in section and partly inelevation, of a resiliently mounted hydrodynamic fluid and preferablygas bearing assembly in accordance with the present invention;

FIG. 1a is a diagrammatic view of the resilient spring mounting assemblyfor the support bearings in accordance with the present invention;

FIG. 2 is a view, partly in section and partly in elevation,illustrating one form of resilient mounting for the support or thrustbearings;

FIG. 3 is a view, partly in section and partly in elevation of anotherform of resilient mount in accordance with this invention;

FIG. 3a is an enlarge fragmentary view of the resilient mountillustrated in FIG. 3;

FIG. 4 is a view, partly in section and partly in elevation,illustrating still another form of resilient mount in accordance withthis invention;

FIG. 5 is a view, partly in section and partly in elevation,illustrating a hydrostatic bearing system in accordance with thisinvention;

FIG. 5a is a view in section taken along the line 5a--5a of FIG. 5;

FIG. 6 is a view, partly in section and partly in elevation of anotherform of resilient mounting in accordance with this invention;

FIG. 7 is a view in perspective of a machined spring for use inaccordance with this invention;

FIG. 8 is a view, partly in section and partly in elevation, of ahydrodynamic bearing in accordance with the present invention in whichconical support bearings are used;

FIG. 8a is an enlarged fragmentary view, partly in section and partly inelevation, of a portion of the resilient bearing support systemillustrated in FIG. 8;

FIG. 8b is an enlarged fragmentary view, partly in section and partly inelevation, of another form of a resilient bearing support system useablewith the structure illustrated in FIG. 8;

FIG. 9 is a view similar to FIG. 8 illustrating a hydrostatic bearing inaccordance with this invention;

FIG. 9a is a section view taken along the line 9a--9a of FIG. 9;

FIG. 10 is a fragmentary view, partly in section and partly inelevation, of a hydrostatic bearing in accordance with this inventionwhich essentially eliminates almost all of the axial frictional forces;

FIG. 10a is a fragmentary sectional view of a hydrostatic bearing inaccordance with this invention using a hollow shell member which ispressurized by a tube;

FIG. 10b is a fragmentary sectional view of a bearing system inaccordance with this invention using a hollow shell with an inlet tubefixed to a manifold;

FIG. 10c is a fragmentary sectional view of another form of bearingsystem in accordance with this invention using a hollow shell whoseinner chamber is sealed by a double-ended self-energizing spring seal;

FIG. 11 is a view, partly in section and partly in elevation, of abearing system in accordance with this invention using conical supportbearings resiliently mounted;

FIG. 12 is a view, partly in section and partly in elevation,illustrating a hydrostatic bearing system using a large and smallconical support bearing resiliently mounted;

FIG. 12a is a sectional view taken along the line 12a--12a of FIG. 12;

FIG. 13 is a view, partly in section and partly in elevation, of ahydrostatic bearing using conical journals and a thrust disc, allresiliently mounted, in accordance with this invention;

FIG. 13a is a view, partly in section and partly in elevation, of abearing system in accordance with this invention using self-adjustingconical support bearings and a thrust plate;

FIG. 14 is a view, partly in section and partly in elevation, of abearing system in accordance with this invention in which the diameterof the journals are larger than the diameter of the shaft;

FIG. 14a is a developed view of a portion of the components illustratedin FIG. 14;

FIG. 15 is an enlarged fragmentary sectional view illustrating a conicalbearing system in accordance with this invention in which an annularmember for providing a thrust point for the machined spring is pinned toa non-rotating member;

FIG. 16 is a view, partly in section and partly in elevation, of abearing system in accordance with this invention in which the conicaljournals are mounted so that the larger ends are outward to provide morestability;

FIG. 16a is a diagrammatic view illustrating the resultant bearingforces of outwardly mounted journals;

FIG. 16b is a view similar to FIG. 16a illustrating the resultantbearing forces for inwardly mounted journals;

FIG. 17 is an enlarged fragmentary sectional view of a bearing system inaccordance with this invention which provides vastly increased thrust;

FIG. 18 is a view partly in section and partly in elevation, of abearing system in accordance with this invention using a smallerstabilizing bearing at one end and a larger bearing at the other end;

FIG. 18a is a sectional view taken along the line 18a--18a of FIG. 18;

FIG. 19 is a view similar to FIG. 18 showing the use of two relativelylarge bearings to produce increased thrust restoring forces inaccordance with this invention;

FIG. 20 is a view partly, in section and partly in elevation, of thedetailed structure of a gas lubricated gas bearing spindle assemblyincorporating the present invention;

FIG. 21 is a plan view of one of the support bearings used in thestructure of the present invention; and

FIG. 22 is a diagrammatic view of a tilter assembly in accordance withthe present invention and which may be used with any of the variousforms described.

DETAILED DESCRIPTION OF THE INVENTION

Referring to drawings which illustrate various forms and preferredembodiments of this invention, FIG. 1 illustrates diagrammatically aresiliently mounted fluid bearing assembly 10 in accordance with thisinvention which may be in the form of a hydrodynamic bearing assembly. Ashaft 12 (which may also be referred to as a rotor) is radiallysupported for rotation by at least two axially spaced resilientlymounted support bearings 16 and 18. The shaft may be rotated by any wellknown means, not shown.

As illustrated, the supporting bearings 16 and 18 are generallycylindrical in shape and surround the shaft 12. The support bearingsrespectively include an interior cylindrical surface 16a and 18a each ofwhich extends for some axial distance along the shaft and each beingspaced from the outer surface of the shaft so as to form a fluid or gasgap therebetween, as shown at 16c and 18c. The outer surface 16b and 18bof each of the support bearings 16 and 18 are each generally cylindricaland may include an annular radially extending shoulder as at 16d and18d.

The support bearings 16 and 18 are received within a supportingstructure generally indicated at 20, the latter being provided withgenerally annular recesses 20a and 20b, as shown. The outer surfaceportions 16b/16d and 18b/18d of the bearings are in spaced relation tothe annular recesses 20a and 20b. The bearings 16 and 18 are resilientlymounted and supported in the supporting structure by resilient mountingmeans generally designated 21 and 22 positioned between the outersurface portions of the bearings and the associated recesses in thesupport structure. In the form illustrated, the resilient mounting meansis in the form of a flexible spring seal-type of element 23 to bedescribed. In general, these elements are of the type illustrated inFIG. 1a.

Thus, for example, the spring seal element 23 includes a shell 23ahaving relatively low coefficient of friction inner and outer continuousand radially spaced surface support portions 23b and 23c one of which iscontinuously in contact with the supporting structure 20 and the otherof which is continuously in contact with the bearing, for example 16. Toassure resiliency, a spring 23d may be used to urge the support portionsapart or in opposite radial directions. One of the advantages of thisinvention is that various combinations of springs and support portionsmay be used, with the structure of the spring being designed to producethe desired resilient and flexible mounting of the support bearings, aswill be described.

Again referring to FIG. 1, each support bearing 16 and 18 preferablyincludes two spaced spring elements arranged axially on each side of theshoulders 16d and 18d, as shown, for the purpose of support in an axialdirection relative to the shaft. The pairs of spring elements arearranged such that the surface support portions of each spring elementis in facing relation to the other, with the inner and outer surfacesupport element contacting the support bearing and the housing,respectively, with the spring element 23d positioned as indicated. Theresilient mounting means 21 and 22 are essentially of the same structureas already described.

In the form illustrated in FIG. 1, an end structure 25 is affixed to thehousing at one end by bolts or the like and the other end of the housing20 is affixed to a housing member 26 which in turn is affixed to lowerhousing member 28 through an annular spacing plate 30. In this form theshaft 12 includes a disk-like thrust pad 34 which is preferablyintegrally formed with the shaft although other arrangements may beused. It is preferred that the outer surface of the shaft and the outersurfaces of the thrust plate be coated with a high wear material such aspure chrome, tungsten carbide, titanium nitride or titanium carbide, forexample.

The thrust pad 34 operates as the surface for providing an axial thrustforce capability. In hydrodynamic bearings, the thrust and radialrestoring forces are generated by relatively high speed surfaces movingin close proximity with each other and separated by a gas gap. It willbe seen that the thrust pad 34 includes a gap 34a and 34b on each sidethereof and these gaps are the region in which the restoring pressureforce is generated. The analysis of hydrodynamic force generation hasbeen extensively studied and documented. The hydrodynamic force ismainly determined by computer solutions of the Reynolds equations forany given bearing geometry. Reference is made to Design of FilmBearings, P. R. Trumpler, McMillan Press, 1966. By this invention,resiliently mounted bearings further enhance the stability and thrustload carrying capacity of the bearing system.

Thus, the support bearings 16 and 18 are resiliently mounted, asdescribed, and operate to generate a fluid restoring force radially. Inaddition, the thrust pad 34 cooperates with annular disk-like thrust padbearings 36 and 37, one positioned on each side of the thrust pad 34 andcooperating to form the gaps 34a and 34b. To generate restoring forcesmore effectively in the axial direction, each thrust pad bearing 36 and374 is also resiliently mounted by resilient means 36a and 37a. Theresilient means may be essentially the same as those described withreference to FIG. 1a, with one surface support portion contacting theassociated thrust pad bearing and the other surface contacting thehousing member 26 or the lower housing member 28, as shown. The annularspacing plate 30 functions to prevent contact between the thrust pad 34and the associated thrust bearings 36 and 37 during start up. If theshaft and thrust pad are coated with the materials mentioned, it ispreferred that the thrust pad bearings and support bearings be made of aceramic or carbon graphite type material. In this way close tolerancemay be maintained with out the possibility of galling or surfacedestruction if contact is inadvertently made during start up or duringrunning.

One of the aspects of this invention which produces unexpected resultsis the use of support bearings and thrust bearings which are 360 degreeunits and either surround the shaft or rotor or are in continuous facingrelation to the thrust pad. In large measure the effectiveness of thisinvention is the relatively rapid response to changing dynamicconditions through the essentially three axis movement of the supportingstructure which assures that the gas gap is maintained. This requiresboth low friction in the relatively moveable parts and a relatively lowfriction and continuous response rather than a system which mustovercome coloumb friction.

The support bearing and thrust pad bearing support system areresiliently mounted and are self-adjusting for maintaining the closedistances between the support bearings and shaft, and between the padbearings and pad to take up thrust loads. Through the use of essentiallycontinuous support and thrust bearings, the adjusting movement thereofis uniform in response, as contrasted to separately mounted supportbearings which move essentially independently of each other. Byresiliently mounting the thrust and support bearings, each is free tomove uniformly such that the entire bearing surface is able to assume anew relative position with respect to the opposed surface. The bearingsare thus self-aligning and self-adjusting for maintaining the necessaryclose spacing between the bearing surface and the opposed surface.

Resilient mounting of 360 degree units also provides for dynamicmovement which in turn allows for a damping effect which maintainsstability and prevents shaft whirl. The result is that the bearingsystem rotates about its true mass and the slight amount of frictioninherent in the mounting system, in conjunction with the small amount ofdamping provided by the gas film, is sufficient to maintain a stablesystem. The use of a 360 degree spring allows for provision of thecorrect spring constant. The construction of the spring to have thecorrect spring constant is well known in the art.

While various forms of springs may be used, for example a 360 degreegarter spring which surround the support bearing or is in contact withthe thrust bearing, the preferred form of the resilient mounting is aself-energized seal. Referring to FIG. 2, for example, the spring seal50 includes an outer cover 52 having radially spaced support portions52a and 52b which are independently flexible. Associated with the coveris a continuous helically wound spring element 55 which functionsinitially to keep the cover and the spaced support portions energizedand expanded against the cooperating surfaces such as the supportbearing and the associated housing. The cover may be made of any of anumber of flexible materials such as Teflon or glass filled Teflon,carbon or carbon graphite-filled Teflon and the like. In the formillustrated, the spaced support portions include annularly extendingseal and support ridges 56a and 56b as shown, with corresponding ridgeson the inner surface (not shown), and an inwardly extending lip 56d toassist in maintaining the spring 55 and the cover 52 in the properrelative position.

The spring 55 provides a resilient mount which can be controlled by thedesign of the spring, that is, how many turns are used and the diameterof the wire. Therefore, this type of spring may be designed to optimizevirtually any gas bearing system in accordance with this invention byusing the optimum spring constant for the resilient mounts as determinedby the gas film spring and the various masses which are being supported.It will be apparent that the resilient mount 50 is a commerciallyavailable item, but the size and nature thereof should be selectedhaving in mind the desired spring character needed.

FIGS. 3 and 3a illustrate another form of resilient mounting in the formof spring seal 60 having an outer cover 62 which includes radiallyspaced support portions 62a and 62b which are independently flexible.Associated with the outer cover 62 is a continuous spring 65 which is inthe form of a cantilever spring. The spring 65 may be formed by anetching process of a stamping process and then is pre-bent to a U-shapedconfiguration as shown. As seen in FIG. 3a, the spaced support portionsincludes portions 66a and 66b which are thicker in cross-section thanthe remaining portions of the cover. The cover may be of any of thematerials already mentioned.

The form of resilient mounting illustrated in FIG. 4 includes an outercover 70 which may be of any of the materials described. The coverincludes the two radially spaced support portions 72a and 72b and aspring element in the form of a continuous flat strip coil spring 74. Itwill be apparent that other forms of resilient mounting may be used inaccordance with the present invention, but the forms illustrated arepreferred forms.

FIG. 5 illustrates diagrammatically a resiliently mounted fluid bearingassembly 80 in accordance with this invention which may be in the formof a hydrostatic bearing assembly. A shaft 83 is radially supported forrotation by at least two axially spaced resiliently mounted supportbearings 83 and 84. The shaft may be rotated by any well known means,not shown.

As illustrated, the supporting bearings 83 and 84 are generallycylindrical in shape and surround the shaft 82. The support bearingsrespectively include an interior cylindrical surface 83a and 84a each ofwhich extends for some axial distance along the shaft and each beingspaced from the outer surface of the shaft so as to form a fluid or gasgap therebetween, as shown at 83b and 84b. The outer surface 83c and 84cas each of the support bearings 83 and 84 are each generally cylindricaland may include an annular radially extending shoulder as at 83d and84d.

The support bearings 83 and 84 are received within a supportingstructure generally indicated at 87, the latter being provided withgenerally annular recesses 87a and 87b, as shown. The outer surfaceportions 83c/83d and 84c/84d of the bearings are in spaced relation tothe annular recesses 87a and 87b. The bearings 83 and 84 are resilientlymounted and supported in the supporting structure by resilient mountingmeans generally designated 90 and 92 positioned between the outersurface portions of the bearings and the associated recesses in thesupport structure. In the form illustrated, the resilient mounting meansis in the form of a flexible spring seal-type of element 93 as alreadydescribed. The spring seals 93 serve two important functions. Theyresiliently mount the support bearings 83 and 84 and also contain thegas so as to assure that it enters the bearing interspace through atleast on aperture 95.

The radially resilient spring seals 93 are maintained in theirrespective axial positions by annular spring members 96a and 96b, onelocated on each side of the spring seals 93, as shown. These springs 96aand 96b are preferably machined axial springs, disclosed for example inmy pending U.S. application Ser. No. 06/940,948 filed on Dec. 12, 1986and my earlier issued U.S. Pat. No. 4,640,653 issued Feb. 3, 1987, whosedisclosures are incorporated herein by reference. The function of thesprings 96a and 96b is to bias the gas pressure in springs seals 93 andto maintain the spring seal in the proper axial position without causingan undue amount of friction induced by said seal elements on theassociated support bearings. The axial biasing springs 96a and 96b alsoallows for complete axial alignment of the support bearings with respectto the shaft 82.

Each support bearing 83 and 84 preferably includes two spaced springseal elements 93 arranged axially on each side of the shoulders. Thepairs of spring elements are arranged such that the surface supportportions of each spring element is in facing relation to the other, withthe inner and outer surface support element contacting the supportbearing and the housing, respectively. The resilient mounting means areessentially of the same structure as already described.

In the form illustrated in FIG. 5, an end structure 100 is affixed tothe housing at one end by bolts or the like and the other end of thehousing 87 is affixed to a housing member 102 which in turn is affixedto lower housing member 103 through an annular spacing plate 110. Inthis form the shaft 82 includes a disk-like thrust pad 115 which ispreferably integrally formed with the shaft although other arrangementsmay be used. It is preferred that the outer surface of the shaft and theouter surfaces of the thrust plate be coated with a high wear material,as already described.

The thrust pad 115 operates as the surface for providing an axial thrustforce capability. In bearings, the thrust and radial restoring forcesare generated by relatively high speed surfaces moving in closeproximity with each other and separated by a gas gap. It will be seenthat the thrust pad 115 includes a gap 115a and 115b on each sidethereof and these gaps are the region in which the restoring pressureforce is generated.

Thus, the support bearings 83 and 84 are resiliently mounted, asdescribed, and operate to generate a fluid restoring force radially. Inaddition, the thrust pad 115 cooperates with annular disk-like thrustpad bearings 116 and 117, one positioned on each side of the thrust pad115 and cooperating to form the gaps 115a and 115b. To generaterestoring forces more effectively in the axial direction, each thrustpad bearing 116 and 117 is also resiliently mounted by resilient means116a and 117a. The resilient means may be essentially the same as thosepreviously described. The annular spacing plate 110 functions to preventcontact between the thrust pad 115 and the associated thrust bearings116 and 117 during start up. If the shaft and thrust pad are coated withthe materials mentioned, it is preferred that the thrust pad bearingsand support bearings be made of a ceramic or carbon graphite typematerial. In this way close tolerance may be maintained with out thepossibility of galling or surface destruction if contact isinadvertently made during start up or during running.

As seen in FIG. 5, the housing structure is provided with means tointroduce gas into the bearing structure. To this end, the housing isprovided with an inlet 120 which communicates with interior passageways120a and 120b communicating with the apertures 95 and with apertures 95aand 95b in the thrust pad bearings 116 and 117 through passage 120c. Thegas exits through an outlet 122.

The resilient mounted thrust pad bearings provide a stable bearingsystem by providing the means for generating the vibratory motion sothat the necessary amount of damping may be effective. This damping isrepresented by the inherent friction as generated between the seals,their respective housings and the support bearings or thrust bearings.The construction in accordance with this invention allows for excellentalignment between the rotating and non-rotating parts of the bearingsystem. It also allows for slight movement compensations due to anytemperature effects which may tend to alter the dimensionalrelationships of the system.

One of the aspects of this invention which produces unexpected resultsin the use of support bearings and thrust bearings which are 360 degreeunits and either surround the shaft or rotor or are in continuous facingrelation to the thrust pad. In large measure the effectiveness of thisinvention is the relatively rapid response to changing dynamicconditions through the essentially three axis movement of the supportingstructure which assures that the gas gap is maintained. This requiresboth low friction in the relatively moveable parts and a relatively lowfriction and continuous response rather than a system which mustovercome coloumb friction.

The support bearing and thrust pad bearing support system areresiliently mounted and are self-adjusting for maintaining the closedistances between the support bearings and shaft, and between the padbearings and pad to take up thrust loads. Through the use of essentiallycontinuous support and thrust bearings, the adjusting movement thereofis uniform in response, as contrasted to separately mounted supportbearings which move essentially independently of each other. Byresiliently mounting the thrust and support bearings, each is free tomove uniformly such that the entire bearing surface is able to assume anew relative position with respect to the opposed surface. The bearingsare thus self-aligning and self-adjusting for maintaining the necessaryclose spacing between the bearing surface and the opposed surface.

Another feature of this form of the invention illustrated in FIG. 5 isthe increase in the desired thrust restoring force of the system, asfollows: If the thrust pad 115 moves to the right, as viewed in thedrawings, a greater pressure force will be generated in the bearing gap115a with respect to pad bearing 116. If the thrust plate moves to theleft, then a greater force will be generated in bearing gap 115b withrespect to pad bearing 117, thus always providing a restoring force. Ineffect as the surfaces move closer together, a higher gap pressureresults since the gas leakage is reduced and more of the high pressuregas supply is maintained in the gap. In the opposite case, if thesurfaces are further apart, there will be greater leakage and lowerpressure and thus lower restoring forces.

In general then, gas bearings in accordance with this invention providea self-restoring force usually in the opposite direction from which anymovement, as between the respective surfaces, has occurred. When aresilient mounting is used, this effect is magnified with respect to thetotal load that any bearing can carry before contact is made between thetwo surfaces defining the gap and the lubricating region.

Referring to FIG. 5a, there are four resilient self-energized face seals117a each defining an area 130 which receives lubricating gas throughpassage 95b. The size of the area 130 is such that in conjunction withthe lubricating gas pressure produces balanced pressure forces acting inopposite directions. These pressure forces are generated between thethrust pad bearing 117 and the thrust plate 115 in the region 115b. Thearea is a function of the bearing system geometry and lubricating fluidpressure. Through the use of resilient mounting of the support and padbearings in accordance with this invention, the resulting bearing systemis dynamically stable and exhibits excellent thrust characteristicswhile being relatively economical to manufacture. The thrust bearingsystem design of this invention is also self-adjusting andself-aligning.

FIG. 6 illustrates another form of resilient mounting in accordance withthis invention. The spring seal 135 includes an outer cover 136 havingthe radial spaced support surfaces 136a and 136b which are flexible, asdescribed. The outer cover is associated with a spring 137 whosefunction is as already described. Here the spring is U-shaped andcontinuous and may be made of various spring characteristics dependingon the gage of metal and the dimensions of the spaces between the ribs138. This type of spring seal is especially useful with the bearingsystem of FIG. 5.

FIG. 7 illustrates a machined spring 140 used as already described. Thespring includes machined grooves 142 forming leaf like members 143 whichtogether form a complete spring member. A machine type of spring may beformed from a variety of different metals and alloys to obtain therequired force versus distance characteristics by control of itsdimensions, as is well known in the art. For example the diameter of andthe thickness of leaf member 143 can be varied to obtain differentcharacteristics. Machined type springs are especially suited for use inthe present invention since the are symmetric and allow almost perfectalignment of the support bearings with respect to the rotating shaft orrotor.

FIG. 8 illustrates a hydrodynamic bearing 150 with conical journals 156and 158 on the ends of a rotor 159. The journals have different halfangles as indicated at 160 thus provided for a greater stiffness on oneend than the other. Such a configuration is preferred in order toprovide sufficient radial load carrying capacity along with the requiredthrust load capacity. By using a differential taper, a greater surfacearea is provided which has a greater radial load carrying capacity. Inthe form illustrated, the surface area of the conical support bearing162 is greater in area than that of the conical support bearing 164 andthis operates to provide a greater radial support. The thrust capabilityof this form is provided by the projected area of the conical supportbearing with respect to the axial direction.

In most cases this projected area should be the same for each journal toassure axial stability of any given bearing system. In some cases, ifthe projected area is not quite the same, different pocketconfigurations or inlet orifices may be used to compensate fordifferences in these projected areas. Since this form is a hydrodynamicbearing system (that is, it generated its own pressure restoring forcesin the areas 172 and 174), the conical configuration and the axialsprings 176 and 178 maintain the optimum spacing in film 172 and 174 forgenerating maximum pressure forces at any given RPM. The axial springs176 and 178 are preferably machined springs of the type described forthe reasons described.

In order to assure maximum dynamic stability, the conical bearings aresupported by resilient support and seal assemblies 180 and 182, eachincluding a double seal member as already described, i.e., to allowradial and angular movement. Gaps 172 and 174 are formed by the spacebetween the conical journals 158 and the conical support bearing 162 and156 and 164. The machined springs 176 and 178 cooperate with theassociated resilient mounting to provide the axial thrust capability asdescribed.

The conical support bearings 162 and 164 may be made from one piece ormay be manufactured such that good bearing material 162a and 164a iscontained in a metallic housing 162b and 164b, as shown. The bearingmaterial may be any of the materials already described so as to preventgalling if incidental contact is made at high rotary speeds. Coatingssuch as those already described may also be used. As previouslydescribed the resilient mounting means and axial springs for the conicalsupport bearings are in turn supported by a housing 190 which in turn isaffixed to end structures 192 and 193 each having circular and radiallyinwardly extending flanges 192a and 193a, each spaced from the end faceof the associated bearing so that the latter may move in the mannerindicated and as needed. The end structures are affixed to the housingby bolts as shown.

Referring to FIGS. 8 and 8a, the axial or thrust loads arecounterbalanced by the force transmitted through annular oppositely andoutwardly facing projections 195 and 196 which extend axially and whichare provided on each of the annular shoulders 197 and 198 of themetallic housings 162b and 164b, if such are used to contain the bearingmaterial. If the support bearings are made of one piece the shouldersmay be formed on the support bearings. The force is counteracted andmaintained by the axially disposed machined springs 176 and 178 so as tomaintain the support bearings 162 and 164, as well as the conicaljournals 156 and 158 in the correct relative spacing with respect to thehousing.

As shown in FIG. 8a, the support bearing suspension system includesresiliently and radially supported spring seal assemblies as alreadydescribed. The spring seals are composed of the inner spring element 23and shell element 23a as previously described in connection with FIG. 1aor the remaining spring seals already described. The spring seals allowfor axial motion of the associated support bearing. The thrust restoringforce, which is the reaction to force F, is transmitted to the housing190 via the annular projection 195 which is transmitted from the spring23 through the shell 23a and then to the axial spring 178 to the housing190 as indicated at B. Essentially the same action takes place in theother support bearing structure, but in an opposite direction.

FIG. 8b illustrates a modification of the structure shown in FIGS. 8 and8a in which the spring seal assembly 200 may be of the type illustratedin FIG. 2 or 3a in that the spring element 201 may be similar to 55 or65. The outer cover 204 may be as previously described. In this form theannular axially extending projection 205 on the shoulder 206 of thesupport bearing 164 contacts the inside face 201a of the spring asindicated. The support bearing 164 transmits the thrust load F throughthe projection 205 and simultaneously through the spring 201 and cover204 into the axial machined spring 178.

One of the important aspects of this invention is that the supportbearings are fee to move axially in order to compensate for thermalexpansion, lubricating fluid pressure changes and thrust load changes.The use of axially disposed machined springs 178 to bias the axial sealforces provides a system which has a minimum of axial frictional forces.This is important since the axial friction force magnitude effectivelydetermines the repeatability and consistency of maintaining the desiredoptimum support bearing and shaft journal bearing clearances. It is forthis reason that the shells, e.g. 23a and 204 are fabricated from thematerials having a relatively low coefficient of friction, as alreadydescribed. It should also be noted that most of the relative sealmovement as between the seals and the bearings occur on the outerperipheral surface of the shells. This resilient mounting and lip-typecontact affords a minimum line type of contact so as to further minimizethe axial frictional forces.

The hydrostatic bearing system 210 of FIG. 9 essentially eliminates allaxial frictional forces. In this form conical journals 216 and 218 areprovided on the shaft or rotor 220 and cooperate with axially spacedsupport bearings 225 and 227, there being a space or gap 228 and 229between the cooperating support bearings and journals, as indicated. Thepressurized lubricating fluid, preferably a gas, enters through anaperture 230 provided in the housing 233, and exits at 234 and flowsthrough the housing channels 236 and 237 and into the spaces or gaps 228and 229. In this form, the support bearings are provided with apertures241 and 243, as shown, which communicate with the channels 236 and 237.

The pressurized fluid is prevented from leaking by self-energized springseals 252, 253, 254 and 256, each of a type already described. Thesespring seals also provide a resilient radial support mounting for theassociated support bearings 225 and 227. The pressurized fluid tends tourge the spring seals apart and are retained by axial spring members261, 262, 263, and 264 which are in the form of machined springs, asalready described. These machines springs, which act as retainingsprings, assist in providing the proper clearance in the gaps 228 and229 while providing an almost perfect alignment between the supportbearings and the journals. Since the springs are machines springs, theiraction is essentially uniform as contrasted to a coil spring. Sincethese springs are flexible along their length, they allow the individualcomponents to align themselves so as to balance the pressure forces.

Axial thrust load capability is provided by the conical design of thejournal and support bearings and the axial thrust load is counteractedby the axial thrust springs 261-264. The axial load is transmitted fromthe support bearings 225 and 227 through an annular ring 268 which isintegral with the associated support bearing and which faces the seal252 and another ring 269, similar to 268, which faces the seal 256. Thespring seals are retained by the machined springs which in turn arerestrained by the housing 233. The axial retaining springs also allowfor compensation in response to shock loads. The annular rings includeannular axially extending projections 271 and 272 through which theaxial loads are transmitted, as already described in connection with theprior figures.

Since a resiliently mounted bearing is inherently more stable than anon-resiliently mounted system, pocketed support bearings may be used.Orifice-fed pocket support bearings produce considerable more restoringforce than the same support bearing without pockets, since the pocketstructure produces greater pressure forces. However, these pocketbearings are much less stable and tend to produce unwanted vibrations.These vibrations are known by various names such a pneumatic hammer andself-excited resonance. If resilient mounted bearings are used,pocket-type bearings can be used and still maintain stable systems, thusproducing a better overall bearing assembly.

The reason that pneumatic hammer occurs is due to a finite time constantassociated with gas lubricated bearings. That is, with gas bearings ittakes time for the bearing to adjust to a change in shaft position andthis lagging time leads to unstable vibrating systems. This instabilityoccurs in bearings in which the gas if fed through a small feed holeinto a pocket of relatively large volume. Under conditions of staticequilibrium, the flow through the feed holes 241 and 243 is equal to theflow out of pockets 275 and 276, see also FIG. 9a associated with eachof the support bearings. In the form illustrated there are three feedholes and three spaced and arcuately shaped pockets in each supportbearing, as seen in FIG. 9a.

The pocket pressure is controlled by the flow through the feed holes andadjusts automatically to changes in the support bearing and journalrelative position due to bearing loads. If, for example, the load isincreased the opposed faces of the support bearing and the journal movecloser together, i.e., the gap closes, and the flow through the feedhole is reduced. The pressure drop across the feed hole is also reducedso that the pocket pressure rises nearer to the gas supply pressure.However, in order for the pocket pressure to adjust quickly to changesin gap dimension or clearance, the pockets must be of a small volume. Apocket of large volume takes longer to fill and to empty, so that pocketpressure reacts sluggishly to changes in clearance. The changes inpocket pressure can lag to such an extent (a large time constant) that,following a random disturbance form an equilibrium position, anoscillation of increasing amplitude is generated which produced anunstable vibrating system. Resiliently mounting the support bearings, asdescribed in connection with this invention, allows the use of pocketbearings without the above noted instability problems.

Pneumatic hammer is caused by a lagging time constant associated withthe volume of the gas film, while self-excited resonances are caused bythe fluid flow phenomenon such as shock waves with a gas or cavitationwith a fluid. Due to the self-adjusting feature of the structure of thisinvention, the bearing system of this invention can allow for very smallclearances during all phases of bearing operation. These smallclearances tend to keep any fluid flow in a laminar state, which tendsto prevent any shock waves, and minimizes the self-exciting type ofresonances.

FIG. 10 illustrates another form of bearing system in accordance withthis invention in which the support bearing 290 is resiliently mountedby self-energized spring seals 292 and 294 which are contained in aninner bearing housing 296, the latter being supported in the mainhousing 298. The inner bearing housing is moveable axially with respectto the main housing. Associated with the inner housing 296 is a axiallydisposed machined spring 300 which counteracts the thrust load producedby the conical journal 302 on the shaft or rotor 305.

Radial movement is provided by the resilient self-energized spring seals292 and 294 which may be of any of the types already described. Thespring seals are retained in the inner housing by an washer 307 and awasher retainer 308 which may be a snap-ring.

Pressurized fluid enters through an opening in the main housing similarto 230 (FIG. 9) and enters passageway 310 in the main housing to enter atube 315 mounted on and carried by the inner housing 296. The tube, ofwhich there may be several, is sealed to the outer housing by an O-ring317. Fluid then flows through a radial passageway 320 which communicateswith an orifice 324 in the conical support bearing 290. In the formillustrated, there is one tube 315 which feeds an annulus 325 whicheffectively forms a manifold for several passageways for severalradially spaced orifices. Pressurized gas flows from the orifices to theclearance or gap 330 between the support bearing and the journal.

In this form of the present invention, the only axial frictional forcesare those which result from the small inlet tube 315 and the O-ring seal317 and for this reason it is desirable to use the manifold arrangementdescribed with only one inlet tube. The structure of this form resultsin a gas bearing in which the assembly is more tolerant of axialposition variations due to thermal expansion and thrust load changes.The axial friction forces due to the self-energized spring seals areessentially completely eliminated because there is no relative axialmotion of the spring seals with respect to the support bearing. Thecomplete support bearing mounting system, including the support bearing290, the spring seals 292, 294 and the inner housing 296 move as a unitin the axial direction thus eliminating essentially all of the axialfriction forces.

The thrust forces are biases by the axial spring 300 which is containedby end cap 331. In this form of the invention, the axial springs 300 aredesigned specifically to match the axial loads produced by the conicaltype design which in turn is designed to withstand any of the externalthrust forces applied to the total bearing system. Near perfectalignment is maintained by the axial fit of the inner housing 296 withrespect to the main housing 298, the latter including an exit port 332for exit of the pressurized fluid. If desired, a pocket type supportbearing may be used, as previously described.

The basic structure illustrated in FIG. 10 may also be used as ahydrodynamic bearing in which the gas film forces are generated byrotation of the journal within the support bearing. In the hydrodynamictype of bearing, the inlet passages and entry orifices are naturallyeliminated. Various gas film region surface geometry my be used such aspockets with steps or herringbone configurations, in order to obtain themost efficient and a higher load carrying hydrodynamic bearing. Thestructure of FIG. 10 may also be used in conjunction with the type ofbearing illustrated in FIG. 9, one on each end of the shaft. Somodified, the structure allows for free axial adjustment due to the selfcontained inner housing structure shown in FIG. 10 while maintaining theaxial position of the main housing.

The structure illustrated in FIG. 10a is again one which provides for aminimum of axial friction. In this form the support bearing 350 is inthe form of a hollow member formed of bearing material 352, as alreadydescribed, mounted in a housing 354, the support bearing face beingconical as illustrated. The housing includes an annular projection 356and a annular chamber 358 is provided with passageways 359 for flow ofpressurized fluid to the gap 360 formed by the face of the supportbearing and the opposed conical face of the journal 362 which is on theshaft or rotor 365. The support bearing structure is mounted in aresilient support structure composed of a radial spring 367 encased in arelatively low friction shell 368, the mounting permitting radialmovement of the support bearing. The spring and shell may be of the typedescribed.

Pressurized fluid enters through inlet 370 in the housing 372 and flowsto a stationary manifold 374 sealed to the housing by two O-ring seals375. The rear face of the support bearing structure includes a plate 377sealed to the housing and bearing proper and the plate carries a tube380 which protrudes axially into the manifold 374. The tube 380 issealed to the bearing and the manifold by small self energized sealrings 382 which may be as already described.

Cooperating with the resilient spring system is an annular cylindricalspacer 384 which keeps the shell 368 and the spring 367 axiallypositioned while axial machined spring 385 is positioned between theprojection 356 and the manifold 374. The axial spring 385 biases theaxial forces on the support bearing by the gas forces in the gap 360,the annular projection 356 being proportioned to fit closely in thehousing 372 to allow sufficient radial motion to assure proper supportbearing function and at the same time to position the support bearingand journal in the center of the housing. This form of the presentinvention provides for minimal axial friction since the main radialsprings 367-368 are not pressurized. Seal 382 which is pressurized bythe pressurized fluid minimizes the axial friction forces on the supportbearing structures because the seal is comparatively small since theseal 382 only seals the tube 380 which is moveable axially and radiallywith respect to the housing.

In the form illustrated in FIG. 10b, the tube 380 is carried by themanifold 374 while the seal 382 is mounted on the support bearingstructure.

FIG. 10c illustrates another form of bearing system in accordance withthis invention using another form of spring seal arrangement. The samereference numerals will be used as in FIG. 10a for the common parts. Inthis form the chamber 358 includes double ended self-energized springseal assembly 390 in which the relatively low friction shell 391 isexpanded by two springs 393 and 394. The use of a double ended type ofseal eliminates axial forces due to gas pressure within the seal memberand also prevents leakage of bearing fluid out of the bearing shellexcept through orifices 359. In this form, fluid flows into the manifold374, through tube 380 and through passageway 395 and into chamber 358 toexpand the double ended seal. A snap ring 397 is used to retain themember 377 to the bearing body 350.

Seal 382 for the tube 380 is much smaller than seal 390. Axial forcesand motion of the support bearing due to gas film pressure in the gap360 is biased by axial spring 385. The form is similar to thosepreviously described in which a relatively small seal element is used onthe tube 380. Essentially the only axial friction is that caused byfriction between the seal 382 and the tube 380 which is firmly fixed tothe bearing shell 350. The slight relative radial movement requiredbetween the support bearing and the housing is allowed by the radialflexibility of the springs 393 and 394.

The bearing system illustrated in FIG. 11 uses two different sizes ofconical support bearings and journal bearings, the support bearingsbeing resiliently mounted on self-energized spring seals 402, 404, 406and 408 whose structure has already been described. One of the conicalsupport bearings 410 and the associated conical journal 412, mounted onthe shaft 414, is larger than the other conical support bearing 416 andthe associated journal 418. This particular structure provides forgreater thrust capability by the use of a stationary radially disposedthrust pad 420 and a rotating radially disposed thrust surface 422provided on the journal 412, thereby creating a thrust producing region425 between the opposed thrust pad and thrust surface.

The thrust producing region is provided with pressurized fluid, whichenters inlet port 427 positioned in the center housing section 428. Thecenter housing is affixed to an end housing 429 and sealed thereto, asindicated at 429a, while the other end of the housing is affixed to theother end housing 430, through a spacer housing 431, each sealed asindicated at 430a and 431a. Pressurized fluid flows from the inlet 427through passageway 435 into inlet orifices 437 to the thrust producingregion 425. Pressurized fluid also flows through orifices 439 and 440 insupport bearings 410 and 416 and into the associated gaps 442 and 443.The pressurized fluid exits through outlet 445.

Support bearing 410 is resiliently mounted and its axial thrust forcesare biased by axial machined springs 446 and 448 which also allow forthe needed radial movement. Support bearing 416 is also resilientlymounted and machined springs are used, as described. The orientation ofthe bearings is such that the cones are in facing orientation, as shown.While this structure is a hydrostatic bearing, it is understood thatwith the appropriate strength springs and correct surface geometry, itmay be used as a hydrodynamic bearing. The particular structuredescribed offers the advantage of greater thrust capability through theuse of a thrust surface while retaining all of the advantages of aresiliently mounted support bearing and the self-adjusting features ofconical support bearings and journals.

The structure illustrated in FIG. 12 is similar to that of FIG. 11 andthe same reference numerals have been applied to the same parts. Theform illustrated in FIG. 12 has greater side load capacity, as will bedescribed.

In the form of FIG. 12, the conical journals 460 and 462, the latterbeing the smaller and referred to as the "tail" bearing and the formerbeing the "main" bearing, and the associated support bearings 464 and465, are arranged such that the cones open in the same direction. Thetail bearing operates to keep the rotating shaft or rotor 414 alignedwhile the larger bearing resists side loads Fs and thrust loads Ft. Thisform has greater side load capacity since the restoring force Fr acts inthe direction of the conical angle of the larger bearing and has agreater fulcrum than the structure of FIG. 11. A thrust bearing area 470formed between the end face of the main journal bearing 460 and thrustpad 473 provides the required thrust force to stabilize the rotor 414 inits desired transverse position by counteracting the thrust produced bythe main journal bearing 460. Reference is also made to FIG. 12aillustrating the thrust pad area 470, the orifices 474 and associatedpassages 474a. The orifices 474 permit flow of pressurized fluid throughpassages 475 in the end housing 476 which is affixed and sealed to thespacer 431.

In the form of FIG. 12, machined springs 446 and 449 function tomaintain the resilient self-energizing seals 402 and 404 in theircorrect position against the axial spring thrust produced by thepressurized fluid in the gap 441. By retaining the seals 402 and 404 intheir respective axial positions by the machined springs, the result isthat static friction is minimized as between the seal and the mainsupport bearing 464 so that the system is self-adjusting and willmaintain the optimum clearance between the journal 460 and the supportbearing 464 in the gap area 442. The main support bearing includes amachined spring 480 which provides the axial force for maintaining themain support bearing axially resiliently in the desired position.

As noted, the resilient self-energizing seals operate to support thesupport bearings resiliently with respect to the housing, to maintainthe support bearings concentric with respect to the housing and tocontain and direct the pressurized fluid while allowing the bearingsystem to be mobile and thus self-adjusting to keep the desired gapclearance.

In operation, the resilient self-adjusting seals are forced againsttheir respective retaining springs by the pressurized fluid. If there isa fluctuation in fluid pressure, it is desired to have the supportbearing move and adjust so as to maintain the optimum fluid filmthickness in the gap 442. If both seals 402 and 404 are retained by asolid member in place of retaining springs 446 and 447, they would notbe able to move axially with respect to the housing and their staticfriction would adversely affect the axial movement of the supportbearing 464. Due to the static friction between the seals and thesupport bearing, complete freedom of axial movement of the supportbearing would not be possible unless the seals are retained by thesprings 446 and 449.

Since the seals are retained by springs, they are free to move if apressure fluctuation takes place. Both seals move towards each otherwith a decrease in pressure, or both seals move away from each otherwith an increase in pressure. The motion of the seals is always oppositeand equal and thus the static friction of one seal on the supportbearing cancels the static friction of the other seal. In this wayalmost complete freedom of axial movement of motion is obtained withoutthe effects of static friction. In this way, one always obtains thecorrect adjustment to ensure the correct film thickness in the gap 442upon start-up by eliminating the effect of friction due to friction ofthe seals. The main journal bearing axial spring 480 is thus used tobias the gas pressure forces caused by the external axial force Ft onthe bearing 464 and transmitted through the gas film 442.

If the static and dynamic friction of the shell used in the seal shellis essentially the same as is the case with TEFLON or other types ofrelatively low coefficient of friction materials, typicallyfluorocarbons, it is possible to use only one seal retaining spring foreach support bearing. That is, seal retaining spring 446 can be replacedby a solid cylindrical member with only axial spring 449 being used.This form of the invention will also eliminate most of the detrimentaleffects of static friction from affecting the self-adjusting features ofthe support bearing 464. If there is an increase in fluid pressure,support bearing 464 will tend to move to the left as seen in thedrawings. The friction of seal 402 held axially stationary by the solidcylindrical member will tend to retard the motion of the support bearing464. However, seal 404 will move to the left as seen in the drawings dueto the pressure increase, since it is retained by the resilientretaining spring 449. This motion will tend to cause support bearing 464to move to the left, thus counteracting the friction caused bystationary seal member 402. In like manner, the sameself-counter-balancing frictional forces will operate if there is adecrease in the pressure of the pressurized fluid. It should also beapparent that only one retaining spring has to be axially resilient andit has to be the member in the direction in which the support bearing464 has to move if there is an increase in fluid pressure, i.e., axialspring seal 447 which is located at the smaller diameter end of the coneof support bearing 464.

It is also apparent that all of the description referring to the rightsupport bearing 464 and related support structure is applicable to theleft support bearing and support structure.

By way of example and not to be construed as a limitation of the presentinvention, the following specific example illustrates how, by axiallymounting the support bearing 464 with axial support spring 480, thethrust capacity may be substantially increased. The following actuallymeasured bearing characteristics illustrate the almost doubling effectof the thrust capacity of a 28 degree cone angle bearing having anoutside diameter of 1.5 inches and using air (non-purified ordehumidified or chilled) at a pressure of 70 PSIG. The maximum thrustload which may be supported by the support bearing 464 in conjunctionwith journal 460 is 80 pounds at zero fluid film thickness in the gap442.

The given gas spring constant is 50,000 pounds per inch which means thataround an operating gap of 0.0004 of an inch, support bearing 464 willexert a force of 60 pounds to the right as viewed in FIG. 12, whilethrust bearing 473 will counter with a force of 60 pounds to the left(0.0004 times 50,000=20 pounds; thus 80 pounds, the maximum force atzero clearance, less 20 pounds=60 pounds). If a thrust of 40 pounds isapplied at Ft, the journal 460 will touch at 442 if spring 480 isreplaced by a solid member. If spring 480 is assumed to be solid, i.e.,not axially resilient, then if a force Ft is applied so that a 0.0004 ofan inch motion of the shaft 414 occurs to the left, the force betweensupport bearing 464 main journal 460 in the gap 442 will be 80 pounds tothe right. Like wise the force occurring to the left by thrust bearing473 will drop from 60 pounds to 40 pounds. Thus the maximum thrust forcecapacity is 80 pounds to the right less 40 pounds to the left, producinga maximum thrust capacity of 40 pounds counteracting Ft. If spring 480is assumed to have a spring constant of 17,000 psi, a greater thrustcapacity is obtained. The explanation is as follows.

Since support bearing 464 has a maximum thrust capacity of 80 pounds inthe equilibrium position with an Ft of zero, the bearing forces are 60pounds. This produces a differential force of 20 pounds. This force inconjunction with the 17,000 psi spring constant will cause a movement of20/17,000=0.001176 of an inch of support bearing to the left as seen inFIG. 12. This movement will cause a decrease of 0.001176×50,000=58.82pounds of thrust between thrust bearing 473 and the face of journal 460in the gap 442. The net thrust to the left due to the fluid forces inthe region 470 is 60-58.82=1.17 pounds. The total thrust capacity istherefore 80-1.17=78.82 pounds to the right to counteract Ft. Thus, byusing a spring at 480, which has a spring constant at 17,000 psi, thethrust capacity of the structure can be increased from 49 pounds toabout 79 pounds, about double as compared to a common non-resilient typeof bearing structure. This added thrust capacity is also obtained in aself-adjusting bearing system which provides the various degrees offreedom already noted.

The form of the invention illustrated in FIG. 12 lends itself to a onepiece solid type of rotor which exhibits little, it any, shifting of thecenter mass during relatively high rotational speeds. Since the tailjournal bearing 462 preferably has a diameter which is essentially thesame as the main section of the rotor or shaft 414, the rotor may beinserted into the center of an electrical motor's stator which may be acomponent of the center housing 428. In this case, the shaft section 414may be the armature of a high speed motor. In this structure, such anarmature may be assembled to the stator without disassembly of therotating member. In the case of this structure, shifting of the centermass of the rotating member at relatively high speeds may be preventedsince the rotor is essentially of one-piece construction, i.e., noassembly or disassembly is required after balancing of the rotatingcomponent.

The form of the present invention illustrated in FIG. 13 is ahydrostatic bearing using similarly dimensioned conical support bearings500 and 502 and matching journals 503 and 504 on shaft 505. The sameparts previously described will have the same reference numerals. Theshaft 505 includes an integral thrust plate 510 located between thejournals having radially extending faces 511 and 512, as illustrated, inwhich axial forces are generated for providing the required thrustbearing capacities. Pressurized fluid enters inlet 427 of the housingsection 428, flows through passages 435 to passages 475 and exitsthrough outlet 445.

A gas gap 513 is provided between journal 503 and associated supportbearing 500, while a gas gap 514 is provided between journal 504 andassociated support bearing 502. Support bearing 502 is resilientlysupported by self-energizes spring assembly 515 and spring assembly 516;support bearing 500 being resiliently supported by seal assemblies 517and 518, each of the same being as previously described. Also providedare axially mounted machined springs 520, 522, 523 and 524, located asindicated in the drawing. The self energized spring seals 515, 516, 517and 518 are free to move axially and are retained by the associatedaxial springs. The axial springs in combination with the conicaljournals and support bearings provide for self-adjustment to maintainthe clearances in the gaps 513 and 514, the required thrust forces beingprovided in the fluid bearing regions 511 and 512. The fluid bearingregions are provided by the rotating thrust plate 510 and the stationarythrust pad 530 and resiliently mounted thrust pad 532, the latterreceiving flow of pressurized gas through orifices 535 from passageway435. The support bearings are also provided with orifices for flow ofgas into the gap.

In the form illustrated in FIG. 13, increased thrust capacity isachieved through the use of a thrust plate 510 in conjunction withstationary bearing surfaces 530 and 532, the latter being resilientlymounted by resilient supported face type self-energized spring seals 534as already described in detail.

The form of bearing assembly illustrated in FIG. 13a uses axially spacedconical support bearings 550 and 552 and associated axially spacedconical journals 553 and 554 on the shaft 555, the shaft having a thrustplate 560 affixed thereto and preferably evenly spaced between thejournals. This form of the invention provides greater thrust capacitythrough the use of the disk-like thrust plate 560 which extends radiallyof the shaft. Associated with the thrust plate are mating thrust platebearing 562 and 564 located on each side of the thrust plate, as shown,and resiliently supported by internally self-energized spring face seals566 and 568 which function as a resilient mounting and as a seal for thepressurized fluid which flows between the thrust plate and the thrustplate bearings. As an alternative, the spring face seals may be replacedby two annular spring face seals, one inwardly open and the otheroutwardly open.

An annular spacer 569 is positioned between housing sections 571 and 572which are sealed together and sealingly assembled to outer sections 573and 574, as shown. The spacer 569 operates to maintain the properspacing in gaps 576 and 579. The resilient mounting of the thrustbearings 562 and 564 allows a greater alignment capability and anincreased thrust capability.

The increased thrust capability may be explained as follows. If theshaft 555 is pushed to the right as seen in the drawing by a force F,the gap 576 closes and a greater pressure force is developed in thatgap. Likewise, gap 579 increases and a lower pressure force will occurin the gap. The difference in the pressure forces and thus the generatedthrust forces will resist the thrust force F. Since the stationarythrust bearings are resiliently mounted, they will also move in the samedirection as the thrust force F. This movement will cause a relativelygreater gap 579, since the resilient face seals 566 and 568 have arelatively low spring constant as compared to the gas film springconstant. This larger gap 579 as compared to a non-resiliently mountedthrust bearing will manifest a lower pressure in the gap thereforeproviding a greater thrust capacity as compared to a fixed,non-resilient mounting. The resiliently mounted thrust bearing systemalso provides for a more stable overall bearing assembly since itsmovement, in association with the dynamic damping, tends to preventpneumatic hammer and the other self-generated oscillations.

Pressurized fluid enters through inlet 584 into ducts 586 and 587 forflow into and through the thrust bearings 562 and 564 and into the gaps576 and 579. Fluid also flows into the gaps 590 and 591 between thesupport bearings and associated journals. The support bearings 550 and552 are resiliently mounted by self-adjusting seal assemblies 595 and596 and their associated machined springs, all of which has already beendescribed.

FIGS. 14 and 14a illustrate a self-adjusting conical bearing system inwhich the conical journal bearing surfaces 600 and 602 are larger indiameter than the main shaft 605 on which they are carried. Thisstructure may be used where greater thrust and radial bearing support isneeded than can be achieved by journals which can fit through the innerdiameter of an associated component such as a motor stator 607. In thisinstance, one of the conical journals, e.g., 602 may be removed from theshaft 605 for assembly purposes. The correct orientation between themoveable journal bearing 602 and the shaft 605 is maintained by pin 610which indexes the respective parts so as to prevent misalignment. Inthis way, out-of-balance conditions in disassembly and assembly areprevented. The conical journal 602 is secured to the shaft 605 by a nut612 and a locking washer 614, as illustrated. The balance of thestructure may be as previously described.

FIG. 15 illustrates a form of the present invention for transmitting thethrust load generated by the conical support bearing 620 and theexternal forces on the bearing assembly through the support bearing tothe housing 623. In this form, the conical support bearing includes aninner housing 625, a portion 625a of which extends radially over the endface 626 of bearing body 630 provided with an inner conical surface, asshown. Connected to the inner housing 625 and the bearing body 630 is anaxially extending cylindrical contact member 632 having a radialshoulder 633. The cylindrical contact member move with the inner housingand bearing body since it is pinned thereto as indicated by pin 635, ofwhich there may be several arranged around the periphery of the contactmember.

The outboard self-energized spring seal 640 is contacted by andcontained by the radial shoulder 633 of the contact member 632, asshown. Thus all of the thrust is directly transmitted to theself-adjusting thrust spring 645, in the form of an annular machinedspring surrounding the shaft 647 and conical journal 650. One end of themachined spring bears against the radial shoulder and the other againstthe housing 623, as shown, Essentially all of the gas load on springseal 640 is counteracted by contact element 632. Machined spring 651transmits the axial thrust gas seal load from seal 640a to the housing623. Thus essentially all of the thrust and seal gas load (all in theaxial direction) is prevented from being transmitted from the supportbearing through the seal 640 to the housing. This form of the inventioneliminates the thrust loads from being transmitted from the bearing 630directly through the self-energizing radial spring 653 which forms partof the seal assembly 640. As shown, pressurized fluid enters throughinlet 655 into the gap 658 between the support bearing and the journaland exits as already described.

FIG. 16 shows a cross-sectional view of a conical bearing system inwhich the conical bearing surfaces 675 and 676, generated by conicaljournals 677 and 678, in conjunction with support bearings shells 680and 682 respectively, are positioned with the large end of said conicalsurfaces outwardly facing from each other. This structure provides themost resistance to side forces such as F2 acting on shaft 684. F3 is therepresentative average bearing resistance force counteracting F2generated by front bearing 677. A smaller reactionary force F4 is therepresentative average force acting on bearing 678. Because cone 677 islarger at the end closest to the imposed side force F2, the resultinggenerated counter torque is greater than in a system whose cones arefacing with the large end towards the middle as previously described.

This just mentioned effect can more easily be understood by comparingFIG. 16a with FIG. 16b, where it can be seen that the required bearingrestoring forces F3 and F4 which are necessary to counteract side loadF2 is much lower in the arrangement of FIG. 16a than that required inthe arrangement of FIG. 16b. This is due to the direction in which theserepresentative forces act and the resulting bearing spacing which occursbetween the large diameter end of the two journal bearings in each ofthese two illustrated cases. FIG. 16a has a much greater distance L thancorresponding bearing distance L2 of FIG. 16b, even though each bearingsystem has the same side load F2, and are of essentially the samebearing sizes. Thus the configuration shown in FIG. 16 is more stableand better suited to side loads.

In order to assemble this bearing system, one journal such as 676 isdetachable from shaft 684. Pin 686 is used to keep journal 676 alignedwith shaft 684 so as to prevent misalignment and out-of-balance forcesupon assembly and reassembly. The support bearings 680 and 682 have thesame construction as that shown and explained in greater detail in FIG.15.

FIG. 17 is a cross section of a bearing system in which conical journal700 has its large end 702 placed outward so as to have the bestresistance to side loads as described in FIGS. 16, 16a, and 16b. In thisembodiment the thrust due to the conical design of journal bearing 700and support bearing 705, produced in bearing gap area 706 iscounterbalanced by pressure from the bearing pressure fluid exhaust ofregion 706 at point 710 into gap region 712 between the end face 713 ofthe journal and the facing surface 714 of the housing 715. The pressurein end face gap region 712 counteracts the thrust produced by theconical bearing in the gap region 706 between the opposed conicalsurfaces. This design is self-balancing as a "stand-alone" one endedsystem.

This embodiment of the present invention illustrates a system which hasconsiderably more thrust capacity against a thrust force F as shown inthe given direction for the following reasons. First, as force F isapplied to the journal bearing 700 and shaft assembly 720, a smallmovement occurs in the direction F. This movement allows an increase inthe control exhaust area 722 thus allowing the pressurized gas in volume712 to exhaust through passage 724. This lowers the gas pressure involume 712, thus causing a greater thrust resisting force opposite to F.At the same time, the force F causes the gap or thickness of region 706to decrease and thus causes less flow from exhaust 710 out of the gap706 to flow into volume 712, thus further decreasing the pressure involume 712, and thereby contributing to the generation of an evergreater thrust resisting force. The third effect of Force F is toincrease the pressure in gap 706 by causing a decrease in the gap filmthickness in this gap. Therefore, in summary, there are three effectswhich contribute to the exceedingly high thrust capabilities: (1)Increase of exhaust area 722. (2) Decrease of fluid flow into volume 712by a decrease of area 710. (3) Increase of film pressure in region 706from fluid entering said region through duct 730 via orifice 736.

Support bearings 705 are supported by self-energizing spring seals 740and 742. Thrust on support bearing 705 is counterbalanced by thrustspring 745 which may be a machined spring. A spacer or spring 746maintains seal 740 in its desired position. Annular member 750 is pinnedto the inner housing 752 of bearing member 705 by pin 754. Bearingmaterial 756 is contained in the inner housing 752.

The form of the present invention illustrated in FIGS. 18 and 18aemploys the structure as described in FIG. 17 at the larger conicalbearing 760 and journal 762. The self-balancing thrust under no externalthrust load conditions is produced by pressurized gasses in region 765exhausting from the gap 766 at point 769. Thrust reaction forcesopposing external force F occur in gaps 776 and 770. The conical journal772 is a radially stabilizing journal bearing and is of such a size asto allow assembly of the rotating unit without disassembly of its parts.That is, the major diameter of the rear bearing journal is of the samesize as the motor armature 775, so as to allow for reassembly withoutdisassembling the rotor unit. The resisting thrust against force F isproduced in conical gap regions 766 and 770. The counter-resistingthrust for keeping the unit in equilibrium occurs in region 765. Theconical bearings are both self-adjusting for maintaining the optimumbearing gap thickness. Bearings 760 and 778 are radially supported byself-energizing spring seals 780, 782, 784, and 786. All thrust loadsare transmitted to housing 788 through springs 790 and 796. Members 794and 798 can be either a spacer or spring to keep sealing springs 780 and784 in their respective required axial positions. Bearing gas or fluidenters through port 803 and is ducted to bearing gap regions 766 and 770via ducts 805 and 807 respectively. The bearings spent gasses exhaustaround front shaft, out of rear, and also through housing at point B.

FIG. 18A is one type of bearing section as seen looking into bearingmember 760 showing bearing inlet orifices 808 and exhausting grooves 810which create separate bearing pads 812. In many cases, just plainconical surfaces with inlet orifices 808 may be used. Various pocketgeometry may be used on bearing member 760, and in conjunction withherringbone geometry on the rotating member in order to optimize andenhance the stability of these systems. It is obvious that if this was apure hydrodynamic bearing, it would generate its own gas film pressuresand operate in the manner as previously described, except no externalpressurized gasses would be required. However, the pressurized systems(hydrostatic) operate at all speeds and in general have greater loadcarrying capacity.

FIG. 19 shows the utilization of two bearing systems as described inFIG. 17, and represented by rear bearing system 850, and front bearingsystem 860. Bearing air enters at 862, and exhausts from ports 864 and866. The exhaust from rear thrust plate 868 and exhaust to the small endof front bearing at point 872, exhaust through ports 864 located infront bearing 867. This bearing configuration resists the thrust loadforce F to a greater extent (by approximately a factor of two) over thesystem described in FIG. 18, since it has approximately twice theprojected area for counteracting thrust forces. The embodiment shown inFIG. 19 is best suited in applications where greater thrust loads arerequired such as for drilling machine spindles. The rear bearing cone876 is secured to shaft 880 by nut 882, and said cone is kept registeredby pin 884.

The resiliently mounted bearing assembly of this invention may be usedin a gas lubricated spindle assembly, as already mentioned. Referring toFIG. 20 which is essentially a one-to-one scale drawing, a hydrostaticgas bearing spindle assembly 1000 is illustrated incorporating thevarious features of the present invention as well as several otherdesirable features, for the purpose of illustrating the advantages ofthis invention. The spindle illustrated is capable of rotational speedsas high as 120,000 RPM or greater, as mentioned, and may be operated aspeeds as low as 15,000 RPM with a thrust load capability of as high as55 to 60 pounds and can withstand side loads of 15 pounds. Asillustrated the unit weighs less than 5 pounds, a factor which issignificant in terms of use as a spindle for circuit board drillingmachines and other precision drilling operations. In general, thisstructure may drive a drill as small as 4 mils in diameter or as largeas 1/4 of an inch. The rotor is driven by an electrical motor, cooled bywater and uses air, filtered through normal shop filters such as 0.05micron filter and which need not be dehumidified. The water need not berefrigerated. Air at 80 psig and at a rate of 2.0 scf/minute is morethan adequate for the spindle and bearing described.

The gas bearing spindle includes a main housing 1002 which is generallycylindrical in shape and to which is affixed an end plug 1003 screwedinto the end of the main housing. The end plug includes a centralaperture 1004 through which a rotatable shaft 1005 extends. Mounted onthe other end is an upper housing assembly 1007 to which is affixed acover assembly 1009, as indicated. The upper housing contains a toolrelease actuating mechanism generally designated 1010, the latter beingconventional and not forming part of this invention. The upper housingalso includes a side extension 1011 to which is gas sealed an electricalconnection terminal 1013 for the wires for the electrical motor.

The electrical wires pass from the side extension into an axialpassageway 1015 in the main housing and are connected to a stator 1017which is supported, such as by a suitable adhesive, in the main housingand bearing against a shoulder 1018, as shown. Mounted on the shaft 1005which extends axially of the main housing is a rotor 1020, the latteroperating to rotate the shaft, there being a small clearance between therotor and the stator. It is apparent that due to the high rotationalspeeds, the rotor should be structured to withstand the same and thestructure to achieve this is well known in the art. The shaft alsocarries two conical journals 1025 and 1030 arranged such that the smallend of the cones are facing in opposite directions and disposed suchthat one journal is on each side of the rotor, the journals rotatingwith the shaft. Journal 1025 is pinned to the shaft and is removabletherefrom for ease of assembly, as already described. Journal 1030 maybe affixed to the shaft by an adhesive or other suitable means. In apreferred form, the journals may be fabricated of hard anodized aluminumand may include a plurality of blind holes 1025a and 1030a on the largediameter ends of the cone to reduce weight. The journals may also befabricated of ceramic materials such as aluminum oxide.

Cooperating with the journals are spaced conical support bearingassemblies 1050 and 1055 which are resiliently mounted in the housingfor the reasons already described. Each of the support bearingassemblies includes a bearing housing 1050a and 1055a which supports anassociated bearing member 1050b and 1055b, the latter having conicalinner surfaces which mate with the contour of the associated journalssuch that in normal operation there is a small gas gap 1050c and 1055cbetween the support bearings and the associated journal. In a preferredform, the bearing material may non-porous graphite and forms with thejournals a non-galling pair of facing surfaces. The outer peripheralsurface of the bearing assemblies each include a radially extendingshoulder 1050d and 1055d, as already described, the latter beingreceived in recesses in the housing as also described. Received withinthe recesses and located on each side of each of the shoulders is aresilient seal assembly 1060, 1062, 1064 and 1066 which may be of any ofthe types already described and which function as already described.

The shoulder of each of the support bearings is provided along its outersurface with a circumferential groove 1070 and 1072, respectively, whicheffectively forms a manifold for flow of air into the gas gap of eachbearing-journal pair. Each support bearing assembly includes a pluralityof apertures 1050e and 1055e which communicate with the groove for flowof air into the gap, and in a preferred form there are four apertures,located at 90 degrees, the diameter of the apertures being about 11.5mils for the reasons already noted.

Referring to FIG. 21, the location of the apertures, for example 1055eis illustrated. Associated with each of the apertures of each of thesupport bearings is a pressure relief groove 1076-1079 provided in theinner face of the bearing material, i.e., the surface facing thejournal. Assuming the rotation of the shaft is as indicated by thearrow, each groove is also off-set radially such that it is closer tothe aperture located in the clockwise location than that in thecounterclockwise location. As shown, the grooves extend from the smallerdiameter end of the cone and terminate short of the larger diameter end,i.e., they do not extend the full axial distance of the support bearingface. The preferred orientation is that the groove and its associatedaperture form an angle of 30 degrees as measured from the center, asillustrated.

These grooves and their location with respect to the apertures and inrelation to the direction of shaft rotation perform two functions whichtend to reduce axial and radial vibrations of the bearing duringoperation. For example, if a portion or all of the gap increases indimension, then the gas pressure drops and the result is that the partstend to move closer together in the region of the pressure drop. As theparts move closer together, the pressure increases as the gap or aportion of the gap becomes smaller. The result is a cycling due tochanges in pressure in the gap which is believed to cause vibrationsbelieved to be due, at least in part, to the time delay required toreach an equilibrium condition. The grooves, it is believed, tend toreduce the time delay, and equilibrium is reached in a much shorterperiod of time thus reducing undesirable vibrations by allowingincreasing gas pressure to bleed off faster.

The second function of the grooves is related to the complex whirlphenomena which tends to cause an epicyclic rotation of the shaft. It isbelieved that this is due to the fact that the gas pressure distributionmay not be uniform in the gap. For example, if the grooves are locatedsymmetrically and evenly between adjacent apertures, the pressurebetween the groove and the associated aperture is higher than thatbetween the aperture and the next adjacent groove. The relatively largeregions of adjacent high and low pressure tends to cause the movementdescribed. By arranging the grooves closer to the associated aperture,the region of comparatively high pressure is reduced and objectionableepicyclic motions are reduced. It has been found that the 30 degreeorientation described operates quite satisfactorily for the structuredescribed. It is apparent that the grooves may be used in any of thestructures already described.

Cooperating with the support bearing assembly 1055 is a machined tilterspring 1080 having a comparatively high resistance to axial compression,as compared to the machined springs already described, and in the formof an annular machined spring illustrated in more detail in FIG. 22. Amachined spring in the form already described may be used although themachined tilter spring 1080 is quite rigid in an axial direction. Itcan, however, tilt with respect to the axis of the shaft as will beapparent from the structure thereof.

Referring to FIG. 22, the machined tilter spring 1080 is composed of twospaced narrow continuous beams 1080a and 1080b with an intermediatecontinuous beam 1080c which is substantially longer in axial dimensionthan beams 1080a and 1080b. Between the beams 1080a and 1080c are twocircumferential slots 1082 and 1083 with a solid section 1084interconnecting beams 1080a and 1080c. There are two solid sections,arranged 180 degrees apart. The lower beam 1080b is separated from beam1080c by circumferential slots 1085, there being two. The latter slots1085 are oriented 180 degrees out of phase with slots 1082 and 1083 andare also separated by two solid sections which are offset 90 degreeswith respect to the solid sections 1084. In effect, the end beams formresiliently bendable beam sections arranged in 90 degree increments, twoof which are located at a 180 degree orientation on each end but offsetby 90 degrees. The result is that the machined tilter spring 1080 isrigid axially but tiltable. The machined tilter spring thus compensatesfor angular movement of the support bearing assembly 1055 with respectto the housing.

Referring again to FIG. 20, one end 1086 of the machined tilter spring1080 bears against the plug 1030 and the other end 1087 bears againstthe bearing housing 1055a. Should the bearing assembly 1055 attempt totilt, the machined tilter spring allows for such movement. Positionedradially outwardly of the machined tilter spring 1080 is a solid sleeve1090 one end of which bears against the spring assembly 1066 and theother end of which bears against the plug 1003, as shown, to maintainsthis seal assembly in position. This sleeve also prevents the bearingassembly 1055 from bottoming out on the plug 1003. Also associated withthe companion seal assembly 1064 is a second solid sleeve 1092, one endof which bears against the seal assembly 1064 and the other end of whichis received in a shoulder 1094 provided on the inner wall of thehousing, as shown.

Cooperating with spring seal assembly 1060 is a machined spring 1096similar to those already described, one end of which bears against theseal assembly 1060 and the other end of which bears against a radiallyinwardly extending section 1097 which forms part of the upper housingassembly. The radial section 1097 includes a plurality of apertures1097a whose function is to permit flow of air, as will be described.Positioned radially inwardly of the spring 1096 is another machinedspring 1099 having a structure as already described, one end of whichbears against the support bearing assembly 1050 and the other end ofwhich bears against the radial section 1097. The machined spring 1099operates as the main axial thrust spring for the bearing assembly whilethe machined spring 1096 cooperates with the associated spring sealassembly to mount the same resiliently for axial movement. Locatedbetween support bearing assembly 1050 and the stator 1017 and bearingagainst an inwardly extending shoulder 1102 on the inner surface of thehousing is a pressure plate 1105 having a radially inwardly extendinglip 1105a and an axially extending lip 1105b, the latter bearing againstthe spring seal assembly 1062. The pressure plate operates to assist indisassembly of the structure by holding the journal 1025 as the shaftand other journal are pushed out of the housing after the end plug 1003is removed.

As illustrated, the shaft 1005 is hollow along its length to receive adrawbar 1110, one end of which is threaded as at 1112 to receive a splitcollet 1114 having an outside tapered surface for holding a tool. Thedetails of the drawbar and collet are as described in my prior patentU.S. Pat. No. 4,640,653 and previously described application. A drivingrotary connection between the drawbar 1110 and the shaft is made by asplined connection 1116 or other suitable means which allows limitedrelative axial movement between the drawbar and the shaft. Threaded onone end of the shaft 1005 is a spring nut 1120 in which the spring is amachined spring having the structure and function of my prior U.S. Pat.No. 4,640,653 and previously identified application whose disclosuresare incorporated herein by reference. The end of the drawbar includes ahead 1123 having an external shoulder 1123a which bears against the endof the spring nut. The end of the spring nut facing the collet is formedwith an internal concave spherical face 1123b which mates with aninternally threaded nut 1125 having one flat face 1125a and one convexspherical face 1125b whose contour matches the concave face 1123b of thespring nut 1123. The flat and planar face 1125 of the nut bears againstthe end face of journal 1025 to hold the latter on the shaft, the latterbeing pinned, as indicated at 1126 to the rotor which is joined to theshaft by adhesives or other suitable means.

The use of mating spherical surfaces between the spring nut and themating nut 1125 offers unique advantages. For example, if care is nottaken in forming the threads on which nut 1125 is mounted, or if the endface of the journal is not formed planar and true with respect to theshaft axis, then as the nut is tightened on the shaft to hold thejournal, a very slight bend may be placed on the shaft which increasesas the temperature increases due to the tendency of the shaft toelongate. If the threads and end face of the journal are formed withextreme care or by accurate machining equipment, conventional nuts maybe used. If there is a small bend in the shaft as a result of threadingproblems, the result tends to be epicyclic motion of the shaft. The useof a nut and spring nut with the mating spherical surfaces eliminatesany such problems by forming essentially a partial ball joint whichprevents shaft bending during assembly or during operation.

As shown, the spring nut 1120 tends to urge the drawbar to the right asseen in the drawing with the result that the collet is in a lockingposition and tightly grips a tool. To release the collet and the tool,rotation of the spindle is stopped and the drawbar is moved to the leftas seen in the drawings. One mechanism by which tool release may beachieved is by a air operated tool release mechanism 1010, the latterbeing conventional in design and not forming part of this invention. Anyof the tool release mechanisms known in the art may be used.

Mechanism 1010 includes a cylinder 1150 received within the end housing1007, the cylinder including an inner radially inwardly shoulder 1152which forms a stop for the axial movement of the spring nut and thedrawbar 1110. The cylinder is sealed to a guide 1154 which receives anO-ring 1154a. Associated with the cylinder 1154 is a head member 1156,moveable axially and sealed to the inner surface of the cylinder by anO-ring, as shown, The cover assembly 1009 includes an air fitting 1157which is O-ring sealed to the head 1156, the latter being securedagainst excessive axial movement by a lock ring 1159 received within thecylinder.

Received within the cylinder is an axially moveable piston 1165, sealedto the cylinder by an O-ring, as shown. The piston includes a centrallylocated shaft 1166 which travels axially in a seat 1170, the latterhaving an inner axial shoulder 1170a and an outer periphery seated on ashoulder 1172 provided on the inner surface of the cylinder. The end1175 of the shaft is in axial alignment with the end of the end of thedrawbar. Located between the seat and the back face of the piston andsurrounding the shaft is a helical spring 1180 which urges the pistonand the shaft away from the end of the drawbar.

When a tool change is desired, rotation is stopped, air is introducedinto the space 1185 between the head and the piston tending to forcethese two elements apart. The piston moves against the spring and theend 1175 of the shaft bears against the end of the drawbar and againstthe spring nut to move the drawbar to the left as seen in the drawing torelease the collet and the tool. After a new tool has been placed in thecollet, the air in chamber 1185 is released and the spring urges thepiston away from the end of the drawbar and the spring nut urges thedrawbar to the right as seen in the drawings to close the collet in theshaft. During rotation, the drawbar rotates with the shaft through thespline connection.

One of the advantages of the present design is that the self balancingnature of the spindle and bearing assembly assures that the tool pointruns on a true axis of rotation. For example, if the collet is notconcentric, the point of the drill travels eccentrically. For largerdiameter drills, e.g., 1/4 of an inch, a small eccentric motion may betolerable, but for small diameter drills, e.g., 4 mils, even smallamounts of eccentric motion may not be tolerable. By the presentinvention, due to the self-balancing nature of the bearing, the systemcompensates for non-concentric collets and tends to keep the point ofthe drill rotating on the true mass center of the drill bit.

To cool the spindle during operation against heat build up due to theheat generated by the electrical motor, the outer surface of the housingis preferable provided with a plurality of passageways 1187 which areinterconnected and extend around the outer surface of the housing. Areverse flow pattern is desired and thus cooling fluid such as water isflowed into the passageways from an inlet, not shown, to a reverse flowpassage 1188 at the front end of the spindle, through the passageways1187 and to an outlet, not shown. The cooling passages are formed by anouter cover 1190 which is affixed to the housing by an adhesive althoughother attachment and fluid tight sealing structures may be used. If anadhesive is used, it is preferred to form the cover with a series ofshoulders such that the diameter to the front of shoulder 1190a isgreater than that of the section to the front of shoulder 1190b. Theouter surface of the housing is correspondingly dimensioned. The use ofshoulders prevents the adhesive from being skived off as the cover isassembled over the housing.

One of the singular advantages of this invention is that all of the gas,e.g., air used to lubricate the bearing, may be caused to exit from thefront of the bearing. This is of practical importance in circuit boarddrilling due to the nature of the cuttings, especially with smalldiameter drills. Air exiting from the front also tends to keep thecollet clear of debris and tends to prevent the tool from being held offcenter due to debris in the collet.

Thus, gas and preferably air under pressure is introduced into thebearing through an inlet 1195 in the housing and flows to a passageway1198 which communicates with grooves 1070 and 1072 which are incommunication with the apertures in the support bearings. Positionedbetween the plug 1003 and the support bearing 1055 and held in placethereby and located in aperture 1004 is a shield 1200 which surroundsbut does not contact the shaft. All of the air which enters the bearingsystem exits between the inside of the shield and the outer surface ofthe shaft.

For example, a portion of the air from the gap between journal 1030 andsupport bearing 1055 exits directly into the space between the shaft andthe shield. Air exiting the other end of the same gap flows into achamber 1202 between the stator and journal 1030 and exits throughapertures 1204 into the exit at the front end. A portion of the air fromthe gap between journal 1025 and support bearing 1050 exits into anannular chamber 1207 and flows between the rotor and stator, partiallycooling the same and then into chamber 1202 and out the passages 1204 tothe exit.

The remaining air from the same gap exits into the annulus radiallyinwardly of spring 1099 and in the region of the spring nut. A portionof that air flows between the drawbar and the shaft and out the exitwhile the remainder flows through apertures 1097a into the region of theside extension 1011, through the aperture 1015 for the wiring and theeither around the outside of the stator or between the rotor and thestator to the exit as already described. The reverse flow of the liquidcoolant tends to cool the air as it leaves the exit. It is thus clearthat all of the incoming air exits out the front of the unit, with theadvantages as noted.

It is to be understood that the features described in connection withthe structure of FIG. 20 may be incorporated in the remaining structuresdescribed.

From the above detailed description of the present invention, it willbecome apparent to those skilled in the art that there are severalsignificant advantages of the present invention. For example there aresingular cost advantages since the construction uses readily availablematerials in a design which is relatively simple to manufacture becauseof the relatively non-critical tolerances and the self-adjusting featureas described. The present structure also avoids the need forsophisticated air purifying and dehumidifying equipment or liquidrefrigeration systems for coolant. It is comparatively light in weightand easy to repair. The amount of air usage is relatively low comparedto currently commercially available unit, already identified.

From a technical performance standpoint, apart from the economies ofmanufacture, use and repair, there are a minimum of air supply holeswhich are relatively large in diameter. The bearing of this invention isnot sensitive to dirt contamination as is the case with currentlyavailable products. The facing surfaces of the bearing and journal areself-polishing rather than self-destructive. In the event of accidentalsurface-to-surface contact, for whatever reason, the opposing surfacestend to lap rather than gall. The bearing system is self-adjusting andautomatically maintains the minimum optimal clearance between theopposed surfaces for proper operation.

The use of conical journals and associated support bearings which areresiliently mounted affords greater thrust and inherent stability. Thegas exhausts from the front end of the assembly to keep debris fromentering the bearing or any tool holder or changer mounted thereto andaids in clearing the workpiece of debris. There is a significantimprovement in the ability of the system to accommodate thrust loads;the use of pockets provides for even greater thrust loads. Temperaturefluctuations seem to have no appreciable affect due to theself-adjusting character of the bearing system. Actual operation of anexperimental unit in accordance with this invention for over 8,000 hourshas demonstrated the reliability and performance of air bearing spindleassemblies in accordance with this invention.

From the above description, a number of various structures has beendescribed and it will become apparent to those skilled in the art thatvarious modifications and alterations may be made by those skilled inthe art in accordance with this disclosure without departing from thespirit and scope of the present invention, as set forth in the appendedclaims.

What is claimed is:
 1. A resiliently mounted fluid bearing system foruse at relatively high rotational speeds comprising:support means, shaftmeans having an axis of rotation and including spaced portions formingspaced journals having outer surface portions and arranged along theaxis thereof, at least two axially spaced bearing means mounted inspaced relation in said support means and surrounding said shaft meansat spaced positions along the axis thereof, each said bearing meansincluding an inner surface portion which is in spaced relation to theouter surface portion of said spaced journals to form a fluid gaptherebetween such that the fluid in said gap supports said shaft forrotation with respect to said spaced bearing means, each said bearingmeans including an outer bearing surface portion in opposed spacedrelation to said support means, each said bearing means includingresilient means cooperating with said support means and surrounding theouter bearing surface of each said bearing means for mounting each saidbearing means resiliently with respect to said support means, at leastone machine spring associated with at least one said resilient means forurging the same axially; means to introduce fluid under pressure intothe fluid gap and for exit of fluid out of said gap; said bearing systembeing part of a spindle assembly; and said shaft including a rotor andsaid support means supporting a stator for effecting rotation of saidrotor.
 2. A resiliently mounted fluid bearing system as set forth inclaim 1 wherein a tool changing assembly is supported by said supportmeans, anddrawbar means having a collet at one end thereof and moveablebetween a tool holding and a tool releasing position by said toolchanging assembly.
 3. A resiliently mounted fluid bearing system as setforth in claim 2 wherein said drawbar means is driven by said shaft forrotation therewith while being moveable axially with respect thereto. 4.A resiliently mounted fluid bearing system as set forth in claim 1wherein said shaft includes means to grip a tool and to effect releasethereof.
 5. A spindle assembly including a resiliently mounted fluidbearing system wherein said spindle assembly is adapted for use atrelatively high rotational speeds comprising:housing means, shaft meanshaving an axis of rotation and including spaced portions forming spacedjournals having outer surface portions and arranged along the axisthereof, at least two axially spaced bearings means mounted in spacedrelation in said support means and surrounding said shaft means atspaced positions along the axis thereof, each said bearing meansincluding an inner surface portion which is in spaced relation to theouter surface portion of said spaced journals to form a fluid gaptherebetween such that the fluid in said gap supports said shaft forrotation with respect to said spaced bearing means, means to effectrotation of said shaft, tool holding and releasing means mounted on saidshaft on one end thereof and rotatable therewith, each said bearingmeans including an outer bearing surface portion in opposed spacedrelation to said housing means, each said bearing means includingresilient means cooperating with said housing means and surrounding theouter bearing surface of each said bearing means for mounting each saidbearing means resiliently with respect to said housing means, each saidresilient means including resiliently biased relatively low frictioninner and outer surface support portions, the inner surface supportportion of said resilient means continuously surrounding and contactingsaid outer cylindrical surface portion of said bearing means, the outersurface support portion of said resilient means continuously contactingsaid housing means, means forming liquid passageways for flow of coolingfluid through said spindle and bearing for cooling thereof, means tointroduce fluid under pressure into the gap between each of saidjournals and associated bearing means, machined spring means associatedwith at least one of said bearing means, and means to effect exit of allof said fluid from said one end thereof.
 6. A structure as set forth inclaim 5 wherein at least one of said resilient means includes at leastone machined spring associated therewith and positioned between saidshaft and said housing means.
 7. A structure as set forth in claim 5wherein said means to effect rotation of said shaft includes a rotormounted on said shaft and a stator supported by said housing means.
 8. Astructure as set forth in claim 5 wherein said shaft is hollow,drawbarmeans mounted in said shaft and rotatable therewith, means to permitaxial movement of said drawbar means with respect to said shaft, saiddrawbar means including a releasable collet at one end thereof, andshield means surrounding said shaft to direct all of the gas exitingsaid gaps out said one end thereof.
 9. A structure as set forth in claim5 wherein said journals and said support bearings are conical in shape.10. A structure as set forth in claim 5 wherein at least one of saidjournals is removable from said shaft.
 11. A structure as set forth inclaim 5 wherein said bearing means includes apertures therein for flowof fluid into said gap,said apertures having a diameter greater than 4mils, pressure relief grooves in said bearing means for flow of fluidout one end of said bearing means, and said pressure relief groovesbeing disposed on the upstream side of said apertures with respect tothe direction of rotation of said shaft and being closer to the adjacentdownstream aperture than to the upstream aperture.
 12. A structure asset forth in claim 11 wherein the angle between the aperture and theassociated pressure relief groove is about 30 degrees.